Method of implementing control logic of compression ignition engine

ABSTRACT

A method of implementing control logic of a compression ignition engine is provided. The engine includes an injector, a variable valve operating mechanism, an ignition plug, at least one sensor, and a processor. The processor outputs the signal to the ignition plug in a specific operating state so that unburnt mixture gas combusts by self ignition after the ignition plug ignites the mixture gas inside a combustion chamber. The method includes determining a geometric compression ratio ε of the engine, and determining control logic defining a valve opening angle CA of an intake valve. The valve opening angle CA (deg) is determined so that the following expression is satisfied, if the geometric compression ratio ε is ε&lt;14, −40ε+800+D≤CA≤60ε−550+D. Here, D is a correction term according to the engine speed NE (rpm), D=3.3×10 −10 NE 3 −1.0×10 −6 NE 2 +7.0×10 −4 NE.

TECHNICAL FIELD

The technology disclosed herein relates to a method of implementingcontrol logic of a compression ignition engine.

BACKGROUND OF THE DISCLOSURE

It is known that combustion by compressed self ignition in which amixture gas combusts instantly without flame propagation maximizes fuelefficiency since the combustion period minimized. However, variousproblems must be solved for automobile engines with regard to thecombustion by compressed self ignition. For example, since the operatingstate and environmental conditions vary greatly in the automotiveapplication, a stable compressed self ignition is a major problem. Thecombustion by compressed self ignition has not yet been put to practicaluse for the automobile engine. In order to solve this problem, forexample, JP4,082,292B2 proposes that an ignition plug ignite the mixturegas, when it is difficult for the compressed self ignition to occurbecause of a low combustion-chamber temperature. By igniting the mixturegas immediately before the compression top dead center, the pressurearound the ignition plug increases to facilitate the compressed selfignition.

Unlike the technology disclosed in JP4,082,292B2 in which the compressedself ignition is assisted by the ignition of the ignition plug, thepresent applicant instead proposes SPCCI (SPark Controlled CompressionIgnition) combustion which is a combination of SI (Spark Ignition)combustion and CI (Compression Ignition) combustion. SI combustion iscombustion accompanied by the flame propagation initiated by forciblyigniting the mixture gas inside the combustion chamber. CI combustion iscombustion initiated by the mixture gas inside the combustion chambercarrying out the compressed self ignition. SPCCI combustion iscombustion in which, when the mixture gas inside the combustion chamberis forcibly ignited to start the combustion by flame propagation, theunburnt mixture gas inside the combustion chamber combusts by thecompression ignition because of a pressure buildup due to the heatgeneration and the flame propagation of the SI combustion. Since SPCCIcombustion includes CI combustion, it is one form of “combustion bycompression ignition.”

CI combustion takes place when the in-cylinder temperature reaches anignition temperature defined by the composition of the mixture gas. Fuelefficiency can be maximized, if the in-cylinder temperature reaches theignition temperature near a compression top dead center and CIcombustion takes place. The in-cylinder temperature increases accordingto the increase in the in-cylinder pressure. The in-cylinder pressure inSPCCI combustion is a result of two pressure buildups: a pressurebuildup by the compression work of the piston in a compression stroke,and a pressure buildup caused by the heat generation of the SIcombustion.

Here, the compression work of the piston is defined by an effectivecompression ratio. If the effective compression ratio is too low, thepressure buildup by the compression work of the piston is small. In thiscase, unless the flame propagation in the SPCCI combustion progressesand the pressure buildup caused by the heat generation of the SIcombustion increases considerably, the in-cylinder temperature cannot beraised to the ignition temperature. As a result, since a small amount ofmixture gas is ignited by the compressed self ignition, and a largeamount of mixture gas combusts by the flame propagation, the combustionperiod becomes longer and fuel efficiency decreases. That is, in orderto stabilize the CI combustion in the SPCCI combustion to maximize fuelefficiency, it is necessary to have the effective compression ratioabove a certain value.

Meanwhile, if CI combustion takes place near a compression top deadcenter because of a high in-cylinder temperature at a compressionstarting timing due to a high ambient temperature, etc., the in-cylinderpressure excessively increases to create excessive combustion noise. Inthis case, the combustion noise can be reduced if the ignition timing isretarded. However, if the ignition timing is retarded, since the CIcombustion takes place when the piston falls considerably in theexpansion stroke, fuel efficiency is lowered. Since the pressure buildupcaused by the heat generation of the SI combustion can be utilized inthe SPCCI combustion, it is effective to lower the effective compressionratio to reduce the pressure buildup by the compression work of thepiston in order to achieve both reduced combustion noise and improvedfuel efficiency. Thus, combustion noise can be kept suitable, withoutlowering fuel efficiency.

In order for a design engineer to put to practical use an engine whichperforms the SPCCI combustion, he/she needs to determine, depending oneach engine operating state, the minimum effective compression ratio atwhich the CI combustion is stabilized, and additionally needs to raisethe effective compression ratio within a permissible combustion noiserange and determine the effective compression ratio without thecombustion noise becoming excessively large. Therefore, the designengineer can put to practical use the engine with maximum fuelefficiency by maximizing the ratio in the CI combustion within the SPCCIcombustion, while suppressing the combustion noise to a tolerable level.

However, since the SPCCI combustion is a new combustion system in theart, no one has found the suitable range for the effective compressionratio until now.

The engine effective compression ratio is determined based on thegeometric compression ratio and a valve opening period (valve openingangle) of an intake valve. When implementing engine control logic forperforming the SPCCI combustion, the design engineer needs to determinea range of the valve opening angle.

Since the maximum in-cylinder pressure increases as the geometriccompression ratio increases, the strength of engine components needs tobe raised, which results in an increase in weight, and an increase inmechanical resistance loss. Meanwhile, in terms of thermal efficiency, alarger expansion ratio which is determined based on the geometriccompression ratio is desirable. Since the maximum in-cylinder pressurevaries by the heat balance based on the combustion mode or the strokecapacity even if the geometric compression ratio is same, the optimalgeometric compression ratio for the new combustion system of SPCCIcombustion has been unknown until now.

SUMMARY OF THE DISCLOSURE

The present inventors diligently examined SPCCI combustion, and as aresult, succeeded in determining a suitable range of a valve openingangle within a range of a geometric compression ratio where the SPCCIcombustion occurs. The present inventors came to invent a method ofimplementing control logic of a compression ignition engine based onthis knowledge.

Specifically, the technology disclosed herein relates to the method ofimplementing the control logic of the compression ignition engine.

The engine includes an injector configured to inject fuel to be suppliedin a combustion chamber, a variable valve operating mechanism configuredto change a valve timing of an intake valve, an ignition plug configuredto ignite a mixture gas inside the combustion chamber, at least onesensor configured to measure a parameter related to an operating stateof the engine, and a processor configured to output a signal to theinjector, the variable valve operating mechanism, and the ignition plug,according to control logic corresponding to the operating state of theengine, in response to the measurement of the at least one sensor.

The processor outputs the signal to the ignition plug in a specificoperating state defined by an engine load and an engine speed so thatunburnt mixture gas combusts by self ignition after the ignition plugignites the mixture gas inside the combustion chamber.

The method includes the steps of determining a geometric compressionratio ε of the engine, and determining the control logic defining avalve opening angle CA of the intake valve. When determining the controllogic, the valve opening angle CA (deg) is determined so that thefollowing expression is satisfied, if the geometric compression ratio εis ε<14,−40ε+800+D≤CA≤60ε−550+D  (a)

where D is a correction term according to the engine speed NE (rpm),D=3.3×10⁻¹⁰ NE ³−1.0×10⁻⁶ NE ²+7.0×10⁻⁴ NE.

The ignition plug ignites the mixture gas inside the combustion chamberin response to the signal from the processor. The combustion starts byflame propagation and then the unburnt mixture gas combusts byself-ignition to complete the combustion. That is, this engine performsSPCCI (Spark Controlled Compression Ignition) combustion.

When implementing the control logic of the engine, the design engineerfirst determines the geometric compression ratio ε of the engine. Whenthe geometric compression ratio is set ε<14, the design engineerdetermines the valve opening angle CA of the intake valve so that theexpression (a) is satisfied. By setting the opening angle so as tosatisfy the expression (a), the engine can keep the combustion noisewithin the allowable range while performing stable CI (compressionignition) combustion in the SPCCI combustion which is a combination ofSI (spark ignition) combustion and the CI combustion, even under variousconditions with different situations of the combustion chamber.Operating the engine for performing the SPCCI combustion according tothe control logic maximizes fuel efficiency.

The index for determining the valve opening angle CA of the intakevalve, which is available when implementing the control logic of theengine for performing the SPCCI combustion has been unknown until now.The design engineer had to repeatedly conduct experiments, etc. undervarious conditions while changing the valve opening angle CA of theintake valve to various timings so as to determine the valve openingangle CA corresponding to the operating state of the engine.

The method of defines the relationship between the geometric compressionratio ε of the engine and the valve opening angle CA of the intake valvein order to achieve suitable SPCCI combustion. The design engineer canput to practical use the engine for performing the SPCCI combustion, bydetermining the valve opening angle CA within the range in which therelationship is satisfied. The design engineer can put to practical usethe engine for performing the SPCCI combustion with very few man-hours.

When implementing the control logic, the valve opening angle CA (deg) isdetermined so that the following expression is satisfied, if thegeometric compression ratio ε is 14≤ε<16.−0.7246ε²+6.7391ε+287.68+D≤CA≤290+D  (b)

Thus, with the engine for performing the SPCCI combustion, fuelefficiency is maximized. In addition, the design engineer can put topractical use the engine with fewer man-hours than the conventional one.

When determining the control logic, the valve opening angle CA (deg) isdetermined so that the following expression is satisfied, if thegeometric compression ratio ε is 16≤ε<16.3.−3.9394ε²+159.53ε−1314.9+D≤CA≤290+D  (c)or−0.7246ε²+6.7391ε+287.68+D≤CA≤0.9096ε²−47.634ε+745.28+D  (d)

Similar to the above, with the engine for performing SPCCI combustion,fuel efficiency is maximized. In addition, the design engineer can putto practical use the engine with fewer man-hours than the conventionalone.

When implementing the control logic, the valve opening angle CA (deg) isdetermined so that the following expression is satisfied, if thegeometric compression ratio ε is 16.3≤ε.−3.9394ε²+159.53ε−1314.9+D≤CA≤290+D  (e)or−12.162ε+403.24+D≤CA≤0.9096ε²−47.634ε+745.28+D  (f)

Similar to the above, with the engine for performing the SPCCIcombustion, fuel efficiency is maximized. In addition, the designengineer can put to practical use the engine with fewer man-hours thanthe conventional one.

When implementing the control logic, the valve opening angle CA (deg) isdetermined so that the following expression is satisfied, if the fuel isa low octane fuel, and if the geometric compression ratio ε is ε<12.7.−40(ε+1.3)+800+D≤CA≤60(ε+1.3)−550+D  (g)

Thus, the engine using the low octane fuel also can maximize fuelefficiency. In addition, the design engineer can put to practical usethe engine with man-hours fewer than the conventional one.

By determining a valve close timing IVC of the intake valve so as tosatisfy both of the expressions (a) and (g) if the geometric compressionratio ε is ε<12.7, the design engineer can set the control logicsuitable for both of the engine using a high octane fuel and the engineusing low octane fuel. Even if the octane number of the fuel isdifferent depending on each destination, the design engineer can set thecontrol logic collectively, which reduces the man-hours.

When implementing the control logic, the valve opening angle CA (deg) isdetermined so that the following expression is satisfied, if the fuel isthe low octane fuel, and if the geometric compression ratio ε is12.7≤ε<14.7.−0.7246(ε+1.3)²+6.7391(ε+1.3)+287.68+D≤CA≤290+D  (h)

When implementing the control logic, the valve opening angle CA (deg) isdetermined so that the following expression is satisfied, if the fuel isthe low octane fuel, and if the geometric compression ratio ε is14.7≤ε<15.−3.9394(ε+1.3)²+159.53(ε+1.3)−1314.9+D≤CA≤290+D  (i)or−0.7246(ε+1.3)²+6.7391(ε+1.3)+287.68+D≤CA≤0.9096(ε+1.3)²−47.634(ε+1.3)+745.28+D  (j)

When implementing the control logic, the valve opening angle CA (deg) isdetermined so that the following expression is satisfied, if the fuel isthe low octane fuel, and if the geometric compression ratio ε is 15≤ε.−3.9394(ε+1.3)²+159.53(ε+1.3)−1314.9+D≤CA≤290+D  (k)or−12.162(ε+1.3)+403.24+D≤CA≤0.9096(ε+1.3)²−47.634(ε+1.3)+745.28+D   (l)

The valve close timing IVC of the intake valve may change as theoperating state of the engine changes, and the valve opening angle CA(deg) is determined for each operating state so as to satisfy theexpressions (a) to (l).

Thus, the engine can stably perform the SPCCI combustion in variousoperating states.

The engine may operate in a low-load operating state where the load is agiven load or lower, according to the control logic.

In general CI combustion, the ignitability degrades when the engine loadis low. In this regard, in the SPCCI combustion, the SI combustion isperformed at the start of the combustion, and the CI combustion startsusing the heat released by the SI combustion. The ignitability of theSPCCI combustion does not degrade even when the engine load is low.

The engine may operate in a minimum load operating state according tothe control logic. That is, the engine may perform the SPCCI combustionin the minimum load operating state.

The geometric compression ratio ε of the engine may be set to satisfy10≤ε<21. Thus, the geometric compression ratio ε can be suitably set.

The engine may have a first mode in which the processor outputs a signalto each of the injector and the variable valve operating mechanism sothat an air-fuel ratio (A/F) that is a weight ratio of air contained inthe mixture gas inside the combustion chamber to the fuel becomes leanerthan a stoichiometric air fuel ratio, and outputs a signal to theignition plug so that the unburnt mixture gas combusts by self ignitionafter the ignition plug ignites the mixture gas inside the combustionchamber, and a second mode in which the processor outputs a signal toeach of the injector and the variable valve operating mechanism so thata gas-fuel ratio (G/F) that is a weight ratio of the entire mixture gasinside the combustion chamber to the fuel becomes leaner than thestoichiometric air fuel ratio, and the A/F becomes the stoichiometricair fuel ratio or richer than the stoichiometric air fuel ratio, andoutputs a signal to the ignition plug so that the unburnt mixture gascombusts by self ignition after the ignition plug ignites the mixturegas inside the combustion chamber.

In the first mode, the engine makes the A/F of the mixture gas leanerthan the stoichiometric air fuel ratio, which improves the fuelefficiency of the engine. On the other hand, in the second mode, theengine makes the G/F leaner than the stoichiometric air fuel ratio andmakes the A/F the stoichiometric air fuel ratio or richer than thestoichiometric air fuel ratio. By the G/F becoming lean, fuel efficiencyimproves. Further, by introducing exhaust gas recirculation (EGR) gasinto the combustion chamber, the combustion stability of the SPCCIcombustion increases.

In such an engine, the valve close timing IVC of the intake valve isdetermined so as to satisfy any one of expressions (a) to (h). Thus, theengine can perform the SPCCI combustion that is a combination of the SIcombustion and the CI combustion stably in both the first mode and thesecond mode.

The engine may be provided with an EGR system configured to introduceburnt gas into the combustion chamber. The processor may output a signalto the EGR system and the ignition plug so that a heat amount ratio usedas an index related to a ratio of an amount of heat generated when themixture gas combusts by flame propagation to the entire amount of heatgenerated when the mixture gas inside the combustion chamber combusts,becomes a target heat amount ratio defined corresponding to theoperating state of the engine.

The heat amount ratio of the SPCCI combustion is less than 100%. Theheat amount ratio of the combustion mode where the combustion completesonly by flame propagation without the combustion by compression ignition(i.e., SI combustion) is 100%.

If the heat amount ratio is increased in the SPCCI combustion, the ratioof the SI combustion increases, which is advantageous in reducingcombustion noise. Whereas, if the heat amount ratio is lowered in theSPCCI combustion, the ratio of the CI combustion increases, which isadvantageous in improving fuel efficiency. The heat amount ratio changesby changing the temperature of the combustion chamber and/or theignition timing. For example, when the temperature inside the combustionchamber is high, the CI combustion starts at an early timing, and theheat amount ratio becomes low. Further, when the ignition timing isadvanced, the SI combustion starts at an early timing, and the heatamount ratio becomes high. By the processor outputting the signal to theEGR system and the ignition plug so that the heat amount ratio becomesthe target heat amount ratio defined corresponding to the operatingstate of the engine, both of the reduction of the combustion noise andthe improvement of the fuel efficiency can be achieved.

The processor may output a signal to the EGR system and the ignitionplug so that the heat amount ratio becomes higher when the load of theengine is higher.

When the engine load increases, the amount of fuel supplied into thecombustion chamber increases and the temperature inside the combustionchamber becomes high. By increasing the heat amount ratio of the SPCCIcombustion when the engine load is high, combustion noise is reduced.

BRIEF DESCRIPTION OF THE DRAWINGS

FIG. 1 is a view illustrating a configuration of an engine.

FIG. 2 is a view illustrating a configuration of a combustion chamber,where an upper portion corresponds to a plan view of the combustionchamber, and a lower portion is a cross-sectional view taken along aline II-II.

FIG. 3 is a plan view illustrating a configuration of the combustionchamber and an intake system.

FIG. 4 is a block diagram illustrating a configuration of an enginecontrol device.

FIG. 5 is a graph illustrating a waveform of SPCCI combustion.

FIG. 6 illustrates maps of the engine, where an upper portion is a mapwhen the engine is warm, a middle figure is a map when the engine ishalf warm, and a lower portion is a map when the engine is cold.

FIG. 7 illustrates the details of the map when the engine is warm.

FIG. 8 illustrates charts of a fuel injection timing, an ignitiontiming, and a combustion waveform in each operating range of the map ofFIG. 7.

FIG. 9 illustrates a layer structure of the engine map.

FIG. 10 is a flowchart illustrating a control process according to alayer selection of the map.

An upper graph of FIG. 11 is a graph illustrating a relationship betweenan engine load and a valve open timing of an intake valve in Layer 2,and a lower graph thereof is a graph illustrating a relationship betweenan engine speed and the valve open timing of the intake valve in Layer2.

An upper graph of FIG. 12 is a graph illustrating a relationship betweenthe engine load and the valve open timing of the intake valve in Layer3, a middle figure thereof is a graph illustrating a relationshipbetween the engine load and a valve close timing of an exhaust valve inLayer 3, and a lower graph thereof is a graph illustrating arelationship between the engine load, and an overlap period of theintake valve and the exhaust valve in Layer 3.

FIG. 13 is a flowchart illustrating a process of an operation control ofthe engine executed by an ECU.

FIG. 14 illustrates a relationship between the engine load and a targetSI ratio.

FIG. 15 is a graph illustrating an occurring range of the SPCCIcombustion versus an EGR rate in Layer 2.

FIG. 16 is one example of a matrix image utilized in order to determinea relationship between a geometric compression ratio and a valve closetiming of the intake valve where the SPCCI combustion is possible inLayer 2.

An upper graph of FIG. 17 illustrates a relationship between thegeometric compression ratio and the valve close timing of the intakevalve where the SPCCI combustion is possible in Layer 2 when a highoctane fuel is used, and a lower graph thereof illustrates arelationship between the geometric compression ratio and the valve closetiming of the intake valve where the SPCCI combustion is possible inLayer 2 when a low octane fuel is used.

FIG. 18 is a graph illustrating a range where the SPCCI combustion isstabilized versus a gas air-fuel ratio (G/F) in Layer 3.

FIG. 19 is one example of a matrix image utilized in order to determinethe relationship between the geometric compression ratio and the valveclose timing of the intake valve where the SPCCI combustion is possiblein Layer 3.

An upper graph of FIG. 20 illustrates a relationship between thegeometric compression ratio and the valve close timing of the intakevalve where the SPCCI combustion is possible in Layer 3 when the highoctane fuel is used, and a lower graph thereof illustrates arelationship between the geometric compression ratio and the valve closetiming of the intake valve where the SPCCI combustion is possible inLayer 3 when the low octane fuel is used.

An upper graph of FIG. 21 illustrates a relationship between thegeometric compression ratio and the valve close timing of the intakevalve where the SPCCI combustion is possible in Layer 2 and Layer 3 whenthe high octane fuel is used, and a lower graph thereof illustrates arelationship between the geometric compression ratio and the valve closetiming of the intake valve where the SPCCI combustion is possible inLayer 2 and Layer 3 when the low octane fuel is used.

FIG. 22 is a flowchart illustrating a procedure of a method ofimplementing control logic of a compression ignition engine.

An upper graph of FIG. 23 illustrates a relationship between thegeometric compression ratio and a valve opening angle of the intakevalve where the SPCCI combustion is possible in Layer 2 when the highoctane fuel is used, and a lower graph thereof illustrates arelationship between the geometric compression ratio and the valveopening angle of the intake valve where the SPCCI combustion is possiblein Layer 2 when the low octane fuel is used.

An upper graph of FIG. 24 illustrates a relationship between thegeometric compression ratio and the valve opening angle of the intakevalve where the SPCCI combustion is possible in Layer 3 when the highoctane fuel is used, and a lower graph thereof illustrates arelationship between the geometric compression ratio and the valveopening angle of the intake valve where the SPCCI combustion is possiblein Layer 3 when the low octane fuel is used.

An upper graph of FIG. 25 illustrates a relationship between thegeometric compression ratio and the valve opening angle of the intakevalve where the SPCCI combustion is possible in Layer 2 and Layer 3 whenthe high octane fuel is used, and a lower graph thereof illustrates arelationship between the geometric compression ratio and the valveopening angle of the intake valve where the SPCCI combustion is possiblein Layer 2 and Layer 3 when the low octane fuel is used.

FIG. 26 is a flowchart illustrating a procedure of the method ofimplementing the control logic of the compression ignition engine.

DETAILED DESCRIPTION OF THE DISCLOSURE

Hereinafter, one embodiment of a method of implementing control logic ofa compression ignition engine will be described in detail with referenceto the accompanying drawings. The following description is one exampleof the engine and the method of implementing the control logic.

FIG. 1 is a view illustrating a configuration of the compressionignition engine. FIG. 2 is a view illustrating a configuration of acombustion chamber of the engine. FIG. 3 is a view illustrating aconfiguration of the combustion chamber and an intake system. Note thatin FIG. 1, an intake side is the left side in the drawing, and anexhaust side is the right side in the drawing. In FIGS. 2 and 3, theintake side is the right side in the drawings, and the exhaust side isthe left side in the drawings. FIG. 4 is a block diagram illustrating aconfiguration of a control device of the engine.

An engine 1 is a four-stroke engine which operates by a combustionchamber 17 repeating an intake stroke, a compression stroke, anexpansion stroke, and an exhaust stroke. The engine 1 is mounted on anautomobile with four wheels. The automobile travels by operating theengine 1. Fuel of the engine 1 is gasoline in this example. The fuel maybe a liquid fuel containing at least gasoline. The fuel may be gasolinecontaining, for example, bioethanol.

(Engine Configuration)

The engine 1 includes a cylinder block 12 and a cylinder head 13 placedthereon. A plurality of cylinders 11 are formed inside the cylinderblock 12. In FIGS. 1 and 2, only one cylinder 11 is illustrated. Theengine 1 is a multi-cylinder engine.

A piston 3 is slidably inserted in each cylinder 11. The pistons 3 areconnected with a crankshaft 15 through respective connecting rods 14.Each piston 3 defines the combustion chamber 17, together with thecylinder 11 and the cylinder head 13. Note that the term “combustionchamber” may be used in a broad sense. That is, the term “combustionchamber” may refer to a space formed by the piston 3, the cylinder 11,and the cylinder head 13, regardless of the position of the piston 3.

As illustrated in the lower part of FIG. 2, a lower surface of thecylinder head 13, i.e., a ceiling surface of the combustion chamber 17,is comprised of a slope 1311 and a slope 1312. The slope 1311 is arising gradient from the intake side toward an injection axial center X2of an injector 6 which will be described later. The slope 1312 is arising gradient from the exhaust side toward the injection axial centerX2. The ceiling surface of the combustion chamber 17 is a so-called“pent-roof” shape.

An upper surface of the piston 3 is bulged toward the ceiling surface ofthe combustion chamber 17. A cavity 31 is formed in the upper surface ofthe piston 3. The cavity 31 is a dent in the upper surface of the piston3. The cavity 31 has a shallow pan shape in this example. The center ofthe cavity 31 is offset at the exhaust side with respect to a centeraxis X1 of the cylinder 11.

A geometric compression ratio ε of the engine 1 is set so as to satisfy10≤ε<30, and preferably satisfy 10≤ε<21. The engine 1 which will bedescribed later performs SPCCI (Spark Controlled Compression Ignition)combustion that is a combination of SI (Spark Ignition) combustion andCI (Compression Ignition) combustion in a part of operating ranges. TheSPCCI combustion controls the CI combustion using heat generated and apressure buildup by the SI combustion. The engine 1 is the compressionignition engine. However, in this engine 1, the temperature of thecombustion chamber 17, when the piston 3 is at a compression top deadcenter (i.e., compression end temperature), does not need to beincreased. In the engine 1, the geometric compression ratio can be setcomparatively low. The low geometric compression ratio becomesadvantageous in a reduction of cooling loss and mechanical loss. Forengines using regular gasoline (a low octane fuel of which the octanenumber is about 91), the geometric compression ratio of the engine 1 is14-17, and for those using high octane gasoline (a high octane fuel ofwhich the octane number is about 96), the geometric compression ratio is15-18.

An intake port 18 is formed in the cylinder head 13 for each cylinder11. As illustrated in FIG. 3, each intake port 18 has a first intakeport 181 and a second intake port 182. The intake port 18 communicateswith the corresponding combustion chamber 17. Although the detailedillustration of the intake port 18 is omitted, it is a so-called “tumbleport.” That is, the intake port 18 has such a shape that a tumble flowis formed in the combustion chamber 17.

Each intake valve 21 is disposed in the intake ports 181 and 182. Theintake valves 21 open and close a channel between the combustion chamber17 and the intake port 181 or 182. The intake valves 21 are opened andclosed at given timings by a valve operating mechanism. The valveoperating mechanism may be a variable valve operating mechanism whichvaries the valve timing and/or valve lift. In this example, asillustrated in FIG. 4, the variable valve operating mechanism has anintake-side electric S-VT (Sequential-Valve Timing) 23. The intake-sideelectric S-VT 23 continuously varies a rotation phase of an intake camshaft within a given angle range. The valve open timing and the valveclose timing of the intake valve 21 vary continuously. Note that theelectric S-VT may be replaced with a hydraulic S-VT, as the intake valveoperating mechanism.

An exhaust port 19 is also formed in the cylinder head 13 for eachcylinder 11. As illustrated in FIG. 3, each exhaust port 19 also has afirst exhaust port 191 and a second exhaust port 192. The exhaust port19 communicates with the corresponding combustion chamber 17.

Each exhaust valve 22 is disposed in the exhaust ports 191 and 192. Theexhaust valves 22 open and close a channel between the combustionchamber 17 and the exhaust port 191 or 192. The exhaust valves 22 areopened and closed at a given timing by a valve operating mechanism. Thevalve operating mechanism may be a variable valve operating mechanismwhich varies the valve timing and/or valve lift. In this example, asillustrated in FIG. 4, the variable valve operating mechanism has anexhaust-side electric SVT 24. The exhaust-side electric S-VT 24continuously varies a rotation phase of an exhaust cam shaft within agiven angle range. The valve open timing and the valve close timing ofthe exhaust valve 22 change continuously. Note that the electric S-VTmay be replaced with a hydraulic S-VT, as the exhaust valve operatingmechanism.

The intake-side electric S-VT 23 and the exhaust-side electric S-VT 24adjust length of an overlap period where both the intake valve 21 andthe exhaust valve 22 open. If the length of the overlap period is madelonger, the residual gas in the combustion chamber 17 can be purged.Moreover, by adjusting the length of the overlap period, internal EGR(Exhaust Gas Recirculation) gas can be introduced into the combustionchamber 17. An internal EGR system is comprised of the intake-sideelectric S-VT 23 and the exhaust-side electric S-VT 24. Note that theinternal EGR system may not be comprised of the S-VT.

The injector 6 is attached to the cylinder head 13 for each cylinder 11.Each injector 6 directly injects fuel into the combustion chamber 17.The injector 6 is one example of a fuel injection part. The injector 6is disposed in a valley part of the pent roof where the slope 1311 andthe slope 1312 meet. As illustrated in FIG. 2, the injection axialcenter X2 of the injector 6 is located at the exhaust side of the centeraxis X1 of the cylinder 11. The injection axial center X2 of theinjector 6 is parallel to the center axis X1. The injection axial centerX2 of the injector 6 and the center of the cavity 31 are in agreementwith each other. The injector 6 faces the cavity 31. Note that theinjection axial center X2 of the injector 6 may be in agreement with thecenter axis X1 of the cylinder 11. In such a configuration, theinjection axial center X2 of the injector 6 and the center of the cavity31 may be in agreement with each other.

Although the detailed illustration is omitted, the injector 6 iscomprised of a multi nozzle-port type fuel injection valve having aplurality of nozzle ports. As illustrated by two-dot chain lines in FIG.2, the injector 6 injects the fuel so that the fuel spreads radiallyfrom the center of the combustion chamber 17. The injector 6 has tennozzle ports in this example, and the nozzle port is disposed so as tobe equally spaced in the circumferential direction.

The injectors 6 are connected to a fuel supply system 61. The fuelsupply system 61 includes a fuel tank 63 configured to store fuel, and afuel supply passage 62 which connects the fuel tank 63 to the injector6. In the fuel supply passage 62, a fuel pump 65 and a common rail 64are provided. The fuel pump 65 pumps fuel to the common rail 64. Thefuel pump 65 is a plunger pump driven by the crankshaft 15 in thisexample. The common rail 64 stores fuel pumped from the fuel pump 65 ata high fuel pressure. When the injector 6 is opened, the fuel stored inthe common rail 64 is injected into the combustion chamber 17 from thenozzle ports of the injector 6. The fuel supply system 61 can supplyfuel to the injectors 6 at a high pressure of 30 MPa or higher. Thepressure of fuel supplied to the injector 6 may be changed according tothe operating state of the engine 1. Note that the configuration of thefuel supply system 61 is not limited to the configuration describedabove.

An ignition plug 25 is attached to the cylinder head 13 for eachcylinder 11. The ignition plug 25 forcibly ignites a mixture gas insidethe combustion chamber 17. The ignition plug 25 is disposed at theintake side of the center axis X1 of the cylinder 11 in this example.The ignition plug 25 is located between the two intake ports 181 and 182of each cylinder. The ignition plug 25 is attached to the cylinder head13 so as to incline downwardly toward the center of the combustionchamber 17. As illustrated in FIG. 2, the electrode of the ignition plug25 faces the inside of the combustion chamber 17 and is located near theceiling surface of the combustion chamber 17. Note that the ignitionplug 25 may be disposed at the exhaust side of the center axis X1 of thecylinder 11. Moreover the ignition plug 25 may be disposed on the centeraxis X1 of the cylinder 11.

An intake passage 40 is connected to one side surface of the engine 1.The intake passage 40 communicates with the intake port 18 of eachcylinder 11. Gas introduced into the combustion chamber 17 flows throughthe intake passage 40. An air cleaner 41 is disposed in an upstream endpart of the intake passage 40. The air cleaner 41 filters fresh air. Asurge tank 42 is disposed near the downstream end of the intake passage40. Part of the intake passage 40 downstream of the surge tank 42constitutes independent passages branched from the intake passage 40 foreach cylinder 11. The downstream end of each independent passage isconnected to the intake port 18 of each cylinder 11.

A throttle valve 43 is disposed between the air cleaner 41 and the surgetank 42 in the intake passage 40. The throttle valve 43 adjusts anintroducing amount of the fresh air into the combustion chamber 17 byadjusting an opening of the throttle valve.

A supercharger 44 is also disposed in the intake passage 40, downstreamof the throttle valve 43. The supercharger 44 boosts gas to beintroduced into the combustion chamber 17. In this example, thesupercharger 44 is a mechanical supercharger driven by the engine 1. Themechanical supercharger 44 may be a root, Lysholm, vane, or acentrifugal type.

An electromagnetic clutch 45 is provided between the supercharger 44 andthe engine 1. The electromagnetic clutch 45 transmits a driving forcefrom the engine 1 to the supercharger 44 or disengages the transmissionof the driving force between the supercharger 44 and the engine 1. Aswill be described later, an ECU 10 switches the disengagement andengagement of the electromagnetic clutch 45 to switch the supercharger44 between ON and OFF.

An intercooler 46 is disposed downstream of the supercharger 44 in theintake passage 40. The intercooler 46 cools gas compressed by thesupercharger 44. The intercooler 46 may be of a water cooling type or anoil cooling type, for example.

A bypass passage 47 is connected to the intake passage 40. The bypasspassage 47 connects an upstream part of the supercharger 44 to adownstream part of the intercooler 46 in the intake passage 40 so as tobypass the supercharger 44 and the intercooler 46. An air bypass valve48 is disposed in the bypass passage 47. The air bypass valve 48 adjustsa flow rate of gas flowing in the bypass passage 47.

The ECU 10 fully opens the air bypass valve 48 when the supercharger 44is turned OFF (i.e., when the electromagnetic clutch 45 is disengaged).The gas flowing through the intake passage 40 bypasses the supercharger44 and is introduced into the combustion chamber 17 of the engine 1. Theengine 1 operates in a non-supercharged state, i.e., a naturalaspiration state.

When the supercharger 44 is turned ON, the engine 1 operates in asupercharged state. The ECU 10 adjusts an opening of the air bypassvalve 48 when the supercharger 44 is turned ON (i.e., when theelectromagnetic clutch 45 is engaged). A portion of the gas which passedthrough the supercharger 44 flows back toward upstream of thesupercharger 44 through the bypass passage 47. When the ECU 10 adjuststhe opening of the air bypass valve 48, a supercharging pressure of thegas introduced into the combustion chamber 17 changes. Note that theterm “supercharging” as used herein refers to a situation where thepressure inside the surge tank 42 exceeds an atmospheric pressure, and“non-supercharging” refers to a situation where the pressure inside thesurge tank 42 becomes below the atmospheric pressure.

In this example, a supercharging system 49 is comprised of thesupercharger 44, the bypass passage 47, and the air bypass valve 48.

The engine 1 has a swirl generating part which generates a swirl flowinside the combustion chamber 17. As illustrated in FIG. 3, the swirlgenerating part has a swirl control valve 56 attached to the intakepassage 40. Among a primary passage 401 coupled to the first intake port181 and a secondary passage 402 coupled to the second intake port 182,the swirl control valve 56 is disposed in the secondary passage 402. Theswirl control valve 56 is a opening control valve which is capable ofchoking a cross section of the secondary passage 402. When the openingof the swirl control valve 56 is small, since an intake flow rate of airflowing into the combustion chamber 17 from the first intake port 181 isrelatively large, and an intake flow rate of air flowing into thecombustion chamber 17 from the second intake port 182 is relativelysmall, the swirl flow inside the combustion chamber 17 becomes stronger.On the other hand, when the opening of the swirl control valve 56 islarge, since the intake flow rates of air flowing into the combustionchamber 17 from the first intake port 181 and the second intake port 182become substantially equal, the swirl flow inside the combustion chamber17 becomes weaker. When the swirl control valve 56 is fully opened, theswirl flow will not occur. Note that the swirl flow circulatescounterclockwise in FIG. 3, as illustrated by white arrows (also seewhite arrows in FIG. 2).

An exhaust passage 50 is connected to the other side surface of theengine 1. The exhaust passage 50 communicates with the exhaust port 19of each cylinder 11. The exhaust passage 50 is a passage through whichexhaust gas discharged from the combustion chambers 17 flows. Althoughthe detailed illustration is omitted, an upstream part of the exhaustpassage 50 constitutes independent passages branched from the exhaustpassage 50 for each cylinder 11. The upper end of the independentpassage is connected to the exhaust port 19 of each cylinder 11.

An exhaust gas purification system having a plurality of catalyticconverters is disposed in the exhaust passage 50. Although illustrationis omitted, an upstream catalytic converter is disposed inside an enginebay. The upstream catalytic converter has a three-way catalyst 511 and aGPF (Gasoline Particulate Filter) 512. The downstream catalyticconverter is disposed outside the engine bay. The downstream catalyticconverter has a three-way catalyst 513. Note that the exhaust gaspurification system is not limited to the illustrated configuration. Forexample, the GPF may be omitted. Moreover, the catalytic converter isnot limited to those having the three-way catalyst. Further, the orderof the three-way catalyst and the GPF may suitably be changed.

Between the intake passage 40 and the exhaust passage 50, an EGR passage52 which constitutes an external EGR system is connected. The EGRpassage 52 is a passage for recirculating part of the exhaust gas to theintake passage 40. The upstream end of the EGR passage 52 is connectedbetween the upstream catalytic converter and the downstream catalyticconverter in the exhaust passage 50. The downstream end of the EGRpassage 52 is connected to an upstream part of the supercharger 44 inthe intake passage 40. EGR gas flowing through the EGR passage 52 flowsinto the upstream part of the supercharger 44 in the intake passage 40,without passing through the air bypass valve 48 of the bypass passage47.

An EGR cooler 53 of water cooling type is disposed in the EGR passage52. The EGR cooler 53 cools the exhaust gas. An EGR valve 54 is alsodisposed in the EGR passage 52. The EGR valve 54 adjusts a flow rate ofthe exhaust gas flowing through the EGR passage 52. By adjusting theopening of the EGR valve 54, an amount of the cooled exhaust gas, i.e.,a recirculating amount of external EGR gas can be adjusted.

In this example, an EGR system 55 is comprised of the external EGRsystem and the internal EGR system. The external EGR system can supplyexhaust gas to the combustion chamber 17 that is a lower temperaturethan the internal EGR system.

The control device of the compression ignition engine includes the ECU(Engine Control Unit) 10 for operating the engine 1. The ECU 10 is acontroller based on a well-known microcomputer, and as illustrated inFIG. 4, includes a processor (e.g., a central processing unit (CPU)) 101which executes a computer program, memory 102 which, for example, iscomprised of a RAM (Random Access Memory) and/or a ROM (Read OnlyMemory), and stores the program and data, and an input/output bus 103which inputs and outputs an electrical signal. The ECU 10 is one exampleof a control part.

As illustrated in FIGS. 1 and 4, various kinds of sensors SW1-SW17 areconnected to the ECU 10. The sensors SW1-SW17 output signals to the ECU10. The sensors include the following sensors:

Airflow sensor SW1: Disposed downstream of the air cleaner 41 in theintake passage 40, and measures a flow rate of fresh air flowing throughthe intake passage 40;

First intake-air temperature sensor SW2: Disposed downstream of the aircleaner 41 in the intake passage 40, and measures the temperature offresh air flowing through the intake passage 40;

First pressure sensor SW3: Disposed downstream of the connected positionof the EGR passage 52 in the intake passage 40 and upstream of thesupercharger 44, and measures the pressure of gas flowing into thesupercharger 44;

Second intake-air temperature sensor SW4: Disposed downstream of thesupercharger 44 in the intake passage 40 and upstream of the connectedposition of the bypass passage 47, and measures the temperature of gasflowed out of the supercharger 44;

Second pressure sensor SW5: Attached to the surge tank 42, and measuresthe pressure of gas downstream of the supercharger 44;

Pressure indicating sensors SW6: Attached to the cylinder head 13corresponding to each cylinder 11, and measures the pressure inside eachcombustion chamber 17;

Exhaust temperature sensor SW7: Disposed in the exhaust passage 50, andmeasures the temperature of the exhaust gas discharged from thecombustion chamber 17;

Linear O₂ sensor SW8: Disposed upstream of the upstream catalyticconverter in the exhaust passage 50, and measures the oxygenconcentration of the exhaust gas;

Lambda O₂ sensor SW9: Disposed downstream of the three-way catalyst 511in the upstream catalytic converter, and measures the oxygenconcentration of the exhaust gas;

Water temperature sensor SW10: Attached to the engine 1 and measures thetemperature of coolant;

Crank angle sensor SW11: Attached to the engine 1 and measures therotation angle of the crankshaft 15;

Accelerator opening sensor SW12: Attached to an accelerator pedalmechanism and measures the accelerator opening corresponding to anoperating amount of the accelerator pedal;

Intake cam angle sensor SW13: Attached to the engine 1 and measures therotation angle of an intake cam shaft;

Exhaust cam angle sensor SW14: Attached to the engine 1 and measures therotation angle of an exhaust cam shaft;

EGR pressure difference sensor SW15: Disposed in the EGR passage 52 andmeasures a pressure difference between the upstream and the downstreamof the EGR valve 54;

Fuel pressure sensor SW16: Attached to the common rail 64 of the fuelsupply system 61, and measures the pressure of fuel supplied to theinjector 6; and

Third intake-air temperature sensor SW17: Attached to the surge tank 42,and measures the temperature of gas inside the surge tank 42, i.e., thetemperature of intake air introduced into the combustion chamber 17.

The ECU 10 determines the operating state of the engine 1 based on thesignals of at least one of the sensors SW1-SW17, and calculates acontrol amount of each device according to the control logic definedbeforehand. The control logic is stored in the memory 102. The controllogic includes calculating a target amount and/or the control amount byusing a map stored in the memory 102.

The ECU 10 outputs electrical signals according to the calculatedcontrol amounts to the injectors 6, the ignition plugs 25, theintake-side electric S-VT 23, the exhaust-side electric S-VT 24, thefuel supply system 61, the throttle valve 43, the EGR valve 54, theelectromagnetic clutch 45 of the supercharger 44, the air bypass valve48, and the swirl control valve 56.

For example, the ECU 10 sets a target torque of the engine 1 based onthe signal of the accelerator opening sensor SW12 and the map, anddetermines a target supercharging pressure. The ECU 10 then performs afeedback control for adjusting the opening of the air bypass valve 48based on the target supercharging pressure and the pressure differencebefore and after the supercharger 44 obtained from the signals of thefirst pressure sensor SW3 and the second pressure sensor SW5 so that thesupercharging pressure becomes the target supercharging pressure.

Moreover, the ECU 10 sets a target EGR rate (i.e., a ratio of the EGRgas to the entire gas inside the combustion chamber 17) based on theoperating state of the engine 1 and the map. The ECU 10 then determinesa target EGR gas amount based on the target EGR rate and an inhaled airamount based on the signal of the accelerator opening sensor SW12, andperforms a feedback control for adjusting the opening of the EGR valve54 based on the pressure difference before and after the EGR valve 54obtained from the signal of the EGR pressure difference sensor SW15 sothat the external EGR gas amount introduced into the combustion chamber17 becomes the target EGR gas amount.

Further, the ECU 10 performs an air-fuel ratio feedback control when agiven control condition is satisfied. For example, the ECU 10 adjuststhe fuel injection amount of the injector 6 based on the oxygenconcentration of the exhaust gas which is measured by the linear O₂sensor SW8 and the lambda O₂ sensor SW9 so that the air-fuel ratio ofthe mixture gas becomes a desired value.

Note that the details of other controls of the engine 1 executed by theECU 10 will be described later.

(Concept of SPCCI Combustion)

The engine 1 performs combustion by compressed self ignition under agiven operating state, mainly to improve fuel consumption and emissionperformance. The combustion by self ignition varies largely at thetiming of the self ignition, if the temperature inside the combustionchamber 17 before a compression starts is nonuniform. Thus, the engine 1performs the SPCCI combustion which is a combination of the SIcombustion and the CI combustion.

The SPCCI combustion is combustion in which the ignition plug 25forcibly ignites the mixture gas inside the combustion chamber 17 sothat the mixture gas carries out the SI combustion by flame propagation,and the temperature inside the combustion chamber 17 increases by theheat generation of the SI combustion and the pressure inside thecombustion chamber 17 increases by the flame propagation so that theunburnt mixture gas carries out the CI combustion by self ignition.

By adjusting the heat amount of the SI combustion, the variation in thetemperature inside the combustion chamber 17 before a compression startscan be absorbed. By the ECU 10 adjusting the ignition timing, themixture gas can be self-ignited at a target timing.

In the SPCCI combustion, the heat release of the SI combustion is slowerthan the heat release in the CI combustion. As illustrated in FIG. 5,the waveform of the heat release rate of SI combustion in the SPCCIcombustion is smaller in the rising slope than the waveform in the CIcombustion. In addition, the SI combustion is slower in the pressurefluctuation (dp/dθ) inside the combustion chamber 17 than the CIcombustion.

When the unburnt mixture gas self-ignites after the SI combustion isstarted, the waveform slope of the heat release rate may become steeper.The waveform of the heat release rate may have an inflection point X ata timing of starting the CI combustion.

After the start in the CI combustion, the SI combustion and the CIcombustion are performed in parallel. Since the CI combustion has alarger heat release than SI combustion, the heat release rate becomesrelatively large. However, since the CI combustion is performed after acompression top dead center, the waveform slope of the heat release ratedoes not become too steep. The pressure fluctuation in the CI combustion(dp/dθ) also becomes comparatively slow.

The pressure fluctuation (dp/dθ) can be used as an index representingcombustion noise. As described above, since the SPCCI combustion canreduce the pressure fluctuation (dp/dθ), it is possible to avoidexcessive combustion noise. Combustion noise of the engine 1 can be keptbelow a tolerable level.

The SPCCI combustion is completed when the CI combustion is finished.The CI combustion is shorter in the combustion period than SIcombustion. The end timing of the SPCCI combustion becomes earlier thanthe SI combustion.

The heat release rate waveform of the SPCCI combustion is formed so thata first heat release rate part Q_(SI) formed by the SI combustion and asecond heat release rate part Q_(CI) formed by the CI combustioncontinue in this order.

Here, a SI ratio is defined as a parameter indicative of acharacteristic of the SPCCI combustion. The SI ratio is defined as anindex related to a ratio of the amount of heat generated by the SIcombustion to the entire amount of heat generated by the SPCCIcombustion. The SI ratio is a ratio of heat amount generated by the twodifferent combustion modes. If the SI ratio is high, the ratio of the SIcombustion is high, and if the SI ratio is low, the ratio in the CIcombustion is high. Alternatively, the SI ratio may be defined as aratio of the amount of heat generated by the SI combustion to the amountof heat generated by the CI combustion. That is, in a waveform 801illustrated in FIG. 5,SI ratio=Q _(SI) /Q _(CI).Here,

Q_(SI): Area of SI combustion; and

Q_(CI): Area of CI combustion.

The engine 1 generates a strong swirl flow inside the combustion chamber17 when performing the SPCCI combustion. The term “strong swirl flow”may be defined as a flow having a swirl ratio of four or higher, forexample. The swirl ratio can be defined as a value obtained by dividingan integrated value of intake flow lateral angular velocities by anengine angular velocity, where the intake flow lateral angular velocityis measured for every valve lift, and the measured values are integratedto obtain the integrated value. Although illustration is omitted, theintake flow lateral angular velocity can be obtained based onmeasurement using known rig test equipment.

When the strong swirl flow is generated in the combustion chamber 17,the swirl flow is stronger in an outer circumferential part of thecombustion chamber 17 and is relatively weaker in a central part. By thewhirlpool resulting from a velocity gradient at the boundary between thecentral part and the outer circumferential part, turbulence energybecomes higher in the central part. When the ignition plug 25 ignitesthe mixture gas in the central part, the combustion speed of the SIcombustion becomes higher by the high turbulence energy.

Flame of the SI combustion is carried by the strong swirl flow insidethe combustion chamber 17 and propagates in the circumferentialdirection. The CI combustion is performed from the outer circumferentialpart to the central part in the combustion chamber 17.

When the strong swirl flow is generated in the combustion chamber 17,the SI combustion can fully be performed before the start in the CIcombustion. Thus, the generation of combustion noise can be reduced, andthe variation in the torque between cycles can be reduced.

(Engine Operating Range)

FIGS. 6 and 7 illustrate maps according to the control of the engine 1.The maps are stored in the memory 102 of the ECU 10. The maps includethree kinds of maps, a map 501, a map 502, and a map 503. The ECU 10uses a map selected from the three kinds of maps 501, 502, and 503according to a wall temperature of the combustion chamber 17 and anintake air temperature, in order to control the engine 1. Note that thedetails of the selection of the three kinds of maps 501, 502, and 503will be described later.

The first map 501 is a map when the engine 1 is warm. The second map 502is a map when the engine 1 is half warm. The third map 503 is a map whenthe engine 1 is cold.

The maps 501, 502, and 503 are defined based on the load and the enginespeed of the engine 1. The first map 501 is roughly divided into threeareas depending on the load and the engine speed. For example, the threeareas include a low load area A1, a middle-to-high load area (A2, A3,and A4), and a high speed area A5. The low load area A1 includes idleoperation, and covers areas of a low engine speed and a middle enginespeed. The middle-to-high load area (A2, A3, and A4) are higher in theload than the low load area A1. The high speed area A5 is higher in theengine speed than the low load area A1 and the middle-to-high load area(A2, A3, and A4). The middle-to-high load area (A2, A3, and A4) isdivided into a middle load area A2, a high-load middle-speed area A3where the load is higher than the middle load area A2, and a high-loadlow-speed area A4 where the engine speed is lower than the high-loadmiddle-speed area A3.

The second map 502 is roughly divided into two areas. For example, thetwo areas include a low-to-middle speed area (B1, B2, and B3) and a highspeed area B4 where the engine speed is higher than the low-to-middlespeed area (B1, B2, and B3). The low-to-middle speed area (B1, B2, andB3) is divided into a low-to-middle load area B1 corresponding to thelow load area A1 and the middle load area A2, a high-load middle-speedarea B2, and a high-load low-speed area B3.

The third map 503 has only one area C1, without being divided into aplurality of areas.

Here, the low speed area, the middle speed area, and the high speed areamay be defined by substantially equally dividing the entire operatingrange of the engine 1 into three areas in the engine speed direction. Inthe example of FIGS. 6 and 7, the engine speed is defined to be a lowspeed if the engine speed is lower than the engine speed N1, a highspeed if the engine speed is higher than or equal to the engine speedN2, and a middle speed if the engine speed is higher than or equal tothe engine speed N1 and lower than the engine speed N2. For example, theengine speed N1 may be about 1,200 rpm, and the engine speed N2 may beabout 4,000 rpm.

Moreover, the low load area may be an area including an operating statewith the light load, the high load area may be an area including anoperating state with full load, and the middle load area may be an areabetween the low load area and the high load area. Moreover, the low loadarea, the middle load area, and the high load area may be defined bysubstantially equally dividing the entire operating range of the engine1 into three areas in the load direction.

The maps 501, 502, and 503 in FIG. 6 illustrate the states andcombustion modes of the mixture gas in the respective areas. A map 504in FIG. 7 corresponds to the first map 501, and illustrates the stateand combustion mode of the mixture gas in each area of the map, theopening of the swirl control valve 56 in each area, and a driving areaand a non-driving area of the supercharger 44. The engine 1 performs theSPCCI combustion in the low load area A1, the middle load area A2, thehigh-load middle-speed area A3, the high-load low-speed area A4, thelow-to-middle load area B1, the high-load middle-speed area B2, and thehigh-load low-speed area B3. The engine 1 performs the SI combustion inother areas, specifically, in the high speed area A5, the high speedarea B4, and the area C1.

(Operation of Engine in Each Area)

Below, the operation of the engine 1 in each area of the map 504 in FIG.7 will be described in detail with reference to the fuel injectiontiming and the ignition timing which are illustrated in FIG. 8. Thehorizontal axis in FIG. 8 is a crank angle. Note that reference numerals601, 602, 603, 604, 605, and 606 in FIG. 8 correspond to the operatingstates of the engine 1 indicated by the reference numerals 601, 602,603, 604, 605, and 606 in the map 504 of FIG. 7, respectively.

(Low load Area)

The engine 1 performs the SPCCI combustion when the engine 1 operates inthe low load area A1.

The reference numeral 601 in FIG. 8 indicates fuel injection timings(reference numerals 6011 and 6012), an ignition timing (referencenumeral 6013), and a combustion waveform (i.e., a waveform indicating achange in the heat release rate with respect to the crank angle:reference numeral 6014), when the engine 1 operates in the operatingstate 601 in the low load area A1. The reference numeral 602 indicatesfuel injection timings (reference numerals 6021 and 6022), an ignitiontiming (reference numeral 6023), and a combustion waveform (referencenumeral 6024), when the engine 1 operates in the operating state 602 inthe low load area A1. The reference numeral 603 indicates fuel injectiontimings (reference numerals 6031 and 6032), an ignition timing(reference numeral 6033), and a combustion waveform (reference numeral6034), when the engine 1 operates in the operating state 603 in the lowload area A1. The operating states 601, 602, and 603 have the sameengine speed, but different loads. The operating state 601 has thelowest load (i.e., light load), the operating state 602 has the secondlowest load (i.e., low load), and the operating state 603 has themaximum load among these states.

In order to improve the fuel efficiency of the engine 1, the EGR system55 introduces the EGR gas into the combustion chamber 17. For example,the intake-side electric S-VT 23 and the exhaust-side electric S-VT 24are provided with a positive overlap period where both the intake valve21 and the exhaust valve 22 are opened near an exhaust top dead center.A portion of the exhaust gas discharged from the combustion chamber 17into the intake port 18 and the exhaust port 19 is re-introduced intothe combustion chamber 17. Since the hot exhaust gas is introduced intothe combustion chamber 17, the temperature inside the combustion chamber17 increases. Thus, it becomes advantageous to stabilize the SPCCIcombustion. Note that the intake-side electric S-VT 23 and theexhaust-side electric S-VT 24 may be provided with a negative overlapperiod where both the intake valve 21 and the exhaust valve 22 areclosed.

Moreover, the swirl generating part forms the strong swirl flow insidethe combustion chamber 17. The swirl ratio is four or higher, forexample. The swirl control valve 56 is fully closed or at a givenopening (closed to some extent). As described above, since the intakeport 18 is the tumble port, an inclined swirl flow having a tumblecomponent and a swirl component is formed in the combustion chamber 17.

The injector 6 injects fuel into the combustion chamber 17 a pluralityof times during the intake stroke (reference numerals 6011, 6012, 6021,6022, 6031, and 6032). The mixture gas is stratified by the multiplefuel injections and the swirl flow inside the combustion chamber 17.

The fuel concentration of the mixture gas in the central part of thecombustion chamber 17 is denser or richer than the fuel concentration inthe outer circumferential part. For example, the air-fuel ratio (A/F) ofthe mixture gas in the central part is 20 or higher and 30 or lower, andthe A/F of the mixture gas in the outer circumferential part is 35 orhigher. Note that the value of the A/F is a value when the mixture gasis ignited, and the same applies to the following description. Since theA/F of the mixture gas near the ignition plug 25 is set 20 or higher and30 or lower, generation of raw NO_(x) during the SI combustion can bereduced. Moreover, since the A/F of the mixture gas in the outercircumferential part is set to 35 or higher, the CI combustionstabilizes.

The A/F of the mixture gas is leaner than the stoichiometric air fuelratio throughout the combustion chamber 17 (i.e., excess air ratio λ>1).For example, the A/F of the mixture gas is 30 or higher throughout thecombustion chamber 17. Thus, the generation of raw NO_(x) can be reducedto improve the emission performance.

When the engine load is low (i.e., in the operating state 601), theinjector 6 performs the first injection 6011 in the first half of anintake stroke, and performs the second injection 6012 in the second halfof the intake stroke. The first half of the intake stroke may be a firsthalf of an intake stroke when the intake stroke is equally divided intotwo, and the second half of the intake stroke may be the rest. Moreover,an injection amount ratio of the first injection 6011 to the secondinjection 6012 may be 9:1, for example.

In the operating state 602 where the engine load is higher, the injector6 initiates the second injection 6022 which is performed in the secondhalf of an intake stroke at a timing advanced from the second injection6012 in the operating state 601. By advancing the second injection 6022,the mixture gas inside the combustion chamber 17 becomes morehomogeneous. The injection amount ratio of the first injection 6021 tothe second injection 6022 may be 7:3 to 8:2, for example.

In the operating state 603 where the engine load is even higher, theinjector 6 initiates the second injection 6032 which is performed in thesecond half of an intake stroke at a timing further advanced from thesecond injection 6022 in the operating state 602. By further advancingthe second injection 6032, the mixture gas inside the combustion chamber17 becomes further homogeneous. The injection amount ratio of the firstinjection 6031 to the second injection 6032 may be 6:4, for example.

After the fuel injection is finished, the ignition plug 25 ignites themixture gas in the central part of the combustion chamber 17 at a giventiming before a compression top dead center (reference numerals 6013,6023, and 6033). The ignition timing may be during a final stage of thecompression stroke. The compression stroke may be equally divided intothree, an initial stage, a middle stage, and a final stage, and thisfinale stage may be used as the final stage of the compression strokedescribed above.

As described above, since the mixture gas in the central part has therelatively high fuel concentration, the ignitability improves and the SIcombustion by flame propagation stabilizes. By the SI combustion beingstabilized, the CI combustion begins at a suitable timing. Thus, thecontrollability in the CI combustion improves in the SPCCI combustion.Further, the generation of the combustion noise is reduced. Moreover,since the A/F of the mixture gas is made leaner than the stoichiometricair fuel ratio to perform the SPCCI combustion, the fuel efficiency ofthe engine 1 can be significantly improved. Note that the low load areaA1 corresponds to Layer 3 described later. Layer 3 extends to the lightload operating range and includes a minimum load operating state.

(Middle-to-High Load Area)

When the engine 1 operates in the middle-to-high load area, the engine 1also performs the SPCCI combustion, similar to the low load area.

The reference numeral 604 in FIG. 8 indicates, in the middle-to-highload area, fuel injection timings (reference numerals 6041 and 6042), anignition timing (reference numeral 6043), and a combustion waveform(reference numeral 6044), when the engine 1 operates in the operatingstate 604 in the middle load area A2. The reference numeral 605indicates a fuel injection timing (reference numeral 6051), an ignitiontiming (reference numeral 6052), and a combustion waveform (referencenumeral 6053), when the engine 1 operates in the operating state 605 inthe high-load low-speed area A4.

The EGR system 55 introduces the EGR gas into the combustion chamber 17.For example, the intake-side electric S-VT 23 and the exhaust-sideelectric S-VT 24 are provided with a positive overlap period where boththe intake valve 21 and the exhaust valve 22 are opened near an exhausttop dead center. Internal EGR gas is introduced into the combustionchamber 17. Moreover, the EGR system 55 introduces the exhaust gascooled by the EGR cooler 53 into the combustion chamber 17 through theEGR passage 52. That is, the external EGR gas with a lower temperaturethan the internal EGR gas is introduced into the combustion chamber 17.The external EGR gas adjusts the temperature inside the combustionchamber 17 to a suitable temperature. The EGR system 55 reduces theamount of the EGR gas as the engine load increases. The EGR system 55may not recirculate the EGR gas containing the internal EGR gas and theexternal EGR gas during the full load.

Moreover, in the middle load area A2 and the high-load middle-speed areaA3, the swirl control valve 56 is fully closed or at a given opening(closed to some extent). In the combustion chamber 17, the strong swirlflow with the swirl ratio of four or higher is formed. On the otherhand, in the high-load low-speed area A4, the swirl control valve 56 isopen.

The air-fuel ratio (A/F) of the mixture gas is the stoichiometric airfuel ratio (A/F≈14.7:1) throughout the combustion chamber 17. Since thethree-way catalysts 511 and 513 purify the exhaust gas discharged fromthe combustion chamber 17, the emission performance of the engine 1 isimproved. The A/F of the mixture gas may be set within a purificationwindow of the three-way catalyst. The excess air ratio λ of the mixturegas may be 1.0±0.2. Note that when the engine 1 operates in thehigh-load middle-speed area A3 including the full load (i.e., themaximum load), the A/F of the mixture gas may be set at thestoichiometric air fuel ratio or richer than the stoichiometric air fuelratio (i.e., the excess air ratio λ of the mixture gas is λ<1)throughout the combustion chamber 17.

Since the EGR gas is introduced into the combustion chamber 17, agas-fuel ratio (G/F) which is a weight ratio of the entire gas to thefuel in the combustion chamber 17 becomes leaner than the stoichiometricair fuel ratio. The G/F of the mixture gas may be 18:1 or higher. Thus,generation of a so-called “knock” is avoided. The G/F may be set 18:1 orhigher and 30:1 or lower. Alternatively, the G/F may be set 18:1 orhigher and 50:1 or lower.

When the engine 1 operates in the operating state 604, the injector 6performs a plurality of fuel injections (reference numerals 6041 and6042) during an intake stroke. The injector 6 may perform the firstinjection 6041 in the first half of the intake stroke and the secondinjection 6042 in the second half of the intake stroke.

Moreover, when the engine 1 operates in the operating state 605, theinjector 6 injects fuel in an intake stroke (reference numeral 6051).

The ignition plug 25 ignites the mixture gas at a given timing near acompression top dead center after the fuel is injected (referencenumerals 6043 and 6052). The ignition plug 25 may ignite the mixture gasbefore the compression top dead center when the engine 1 operates in theoperating state 604 (reference numeral 6043). The ignition plug 25 mayignite the mixture gas after the compression top dead center when theengine 1 operates in the operating state 605 (reference numeral 6052).

Since the A/F of the mixture gas is set to the stoichiometric air fuelratio and the SPCCI combustion is performed, the exhaust gas dischargedfrom the combustion chamber 17 can be purified using the three-waycatalysts 511 and 513. Moreover, the fuel efficiency of the engine 1improves by introducing the EGR gas into the combustion chamber 17 andmaking the mixture gas leaner. Note that the middle-to-high load areasA2, A3, and A4 correspond to Layer 2 described later. Layer 2 extends tothe high load area and includes the maximum load operating state.

(Operation of Supercharger)

Here, as illustrated in the map 504 of FIG. 7, the supercharger 44 isOFF in part of the low load area A1 and part of the middle load area A2(see S/C OFF). In detail, the supercharger 44 is OFF in an area on thelower engine speed side in the low load area A1. In an area on thehigher speed side in the low load area A1, the supercharger 44 is ON inorder to secure a required intake filling amount for the increased speedof the engine 1. Moreover, the supercharger 44 is OFF in a partial areaon the lower load and lower engine speed side in the middle load areaA2. In the area on the higher load side in the middle load area A2, thesupercharger 44 is ON in order to secure a required intake fillingamount for the increased fuel injection amount. Moreover, thesupercharger 44 is ON also in the area on the higher speed side in themiddle load area A2.

Note that in each area of the high-load middle-speed area A3, thehigh-load low-speed area A4, and the high speed area A5, thesupercharger 44 is entirely ON (see S/C ON).

(High Speed Area)

When the speed of the engine 1 increases, a time required for changingthe crank angle by 1° becomes shorter. Thus, it becomes difficult tostratify the mixture gas inside the combustion chamber 17. When thespeed of the engine 1 increases, it also becomes difficult to performthe SPCCI combustion.

Therefore, the engine 1 performs not the SPCCI combustion but the SIcombustion when the engine 1 operates in the high speed area A5. Notethat the high speed area A5 extends entirely in the load direction fromthe low load to the high load.

The reference numeral 606 of FIG. 8 indicates a fuel injection timing(reference numeral 6061), an ignition timing (reference numeral 6062),and a combustion waveform (reference numeral 6063), when the engine 1operates in the high speed area A5 in the operating state 606 where theload is high.

The EGR system 55 introduces the EGR gas into the combustion chamber 17.The EGR system 55 reduces the amount of the EGR gas as the loadincreases. The EGR system 55 may not recirculate the EGR gas during fullload.

The swirl control valve 56 is fully opened. No swirl flow occurs in thecombustion chamber 17, but only a tumble flow occurs. By fully openingthe swirl control valve 56, it is possible to increase the fillingefficiency, and reduce the pumping loss.

Fundamentally, the air-fuel ratio (A/F) of the mixture gas is thestoichiometric air fuel ratio (A/F≈14.7:1) throughout the combustionchamber 17. The excess air ratio λ of the mixture gas may be set to1.0±0.2. Note that the excess air ratio λ of the mixture gas may belower than 1 when the engine 1 operates near full load.

The injector 6 starts the fuel injection during an intake stroke. Theinjector 6 injects fuel all at once (reference numeral 6061). Bystarting the fuel injection in the intake stroke, the homogeneous orsubstantially homogeneous mixture gas is formed inside the combustionchamber 17. Moreover, since a longer vaporizing time of the fuel can besecured, unburnt fuel loss can also be reduced.

After the fuel injection is finished, the ignition plug 25 ignites themixture gas at a suitable timing before a compression top dead center(reference numeral 6062).

(Layer Structure of Map)

As illustrated in FIG. 9, the maps 501, 502, and 503 of the engine 1illustrated in FIG. 6 are comprised of a combination of three layers,Layer 1, Layer 2, and Layer 3.

Layer 1 is a layer used as a base layer. Layer 1 extends throughout theoperating range of the engine 1. Layer 1 corresponds to the entire thirdmap 503.

Layer 2 is a layer which is superimposed on Layer 1. Layer 2 correspondsto a portion of the operating range of the engine 1. For example, Layer2 corresponds to the low-to-middle speed area B1, B2, and B3 of thesecond map 502.

Layer 3 is a layer which is superimposed on Layer 2. Layer 3 correspondsto the low load area A1 of the first map 501.

Layer 1, Layer 2, and Layer 3 are selected according to the walltemperature of the combustion chamber 17 and the intake air temperature.

When the wall temperature of the combustion chamber 17 is higher than agiven first wall temperature (e.g., 80° C.) and the intake airtemperature is higher than a given first intake air temperature (e.g.,50° C.), Layer 1, Layer 2, and Layer 3 are selected, and the first map501 is formed by superimposing Layer 1, Layer 2, and Layer 3. In the lowload area A1 in the first map 501, the top Layer 3 therein becomeseffective, in the middle-to-high load areas A2, A3, and A4, the topLayer 2 therein becomes effective, and in the high speed area A5, Layer1 becomes effective.

When the wall temperature of the combustion chamber 17 is lower than thegiven first wall temperature and higher than a given second walltemperature (e.g., 30° C.), and the intake air temperature is lower thanthe given first intake air temperature and higher than a given secondintake air temperature (e.g., 25° C.), Layer 1 and Layer 2 are selected.By piling up the Layer 1 and Layer 2, the second map 502 is formed. Thelow-to-middle speed area B1, B2, and B3 in second map 502, the top Layer2 therein becomes effective, and in the high speed area B4, Layer 1becomes effective.

When the wall temperature of the combustion chamber 17 is lower than thegiven second wall temperature and the intake air temperature is lowerthan the given second intake air temperature, only Layer 1 is selectedto form the third map 503.

Note that the wall temperature of the combustion chamber 17 may bereplaced, for example, by temperature of the coolant of the engine 1measured by the water temperature sensor SW10. Alternatively, the walltemperature of the combustion chamber 17 may be estimated based on thetemperature of the coolant, or other measurements. The intake airtemperature is measurable, for example, by the third intake-airtemperature sensor SW17 which measures the temperature inside the surgetank 42. Alternatively, the temperature of the intake air introducedinto the combustion chamber 17 may be estimated based on various kindsof measurements.

As described above, the SPCCI combustion is performed by generating thestrong swirl flow inside the combustion chamber 17. Since the flamepropagates along the wall of the combustion chamber 17 during the SIcombustion, the flame propagation of the SI combustion is influenced bythe wall temperature. If the wall temperature is low, the flame of theSI combustion is cooled to delay the timing of compression ignition.

Since the CI combustion of the SPCCI combustion is performed in the areafrom the outer circumferential part to the central part of thecombustion chamber 17, it is influenced by the temperature in thecentral part of the combustion chamber 17. If the temperature in thecentral part is low, the CI combustion becomes unstable. The temperaturein the central part of the combustion chamber 17 depends on thetemperature of the intake air introduced into the combustion chamber 17.That is, when the intake air temperature is higher, the temperature inthe central part of the combustion chamber 17 becomes higher, and whenthe intake air temperature is lower, the temperature in the central partbecomes lower.

When the wall temperature of the combustion chamber 17 is lower than thegiven second wall temperature and the intake air temperature is lowerthan the given second intake air temperature, the stable SPCCIcombustion cannot be performed. Thus, only Layer 1 which performs the SIcombustion is selected, and the ECU 10 operates the engine 1 based onthe third map 503. By the engine 1 performing the SI combustion in theentire operating range, the combustion stability can be secured.

When the wall temperature of the combustion chamber 17 is higher thanthe given second wall temperature and the intake air temperature ishigher than the given second intake air temperature, the stable SPCCIcombustion of the mixture gas having substantially stoichiometric airfuel ratio (i.e., λ≈1) can be carried out. Thus, in addition to Layer 1,Layer 2 is selected, and the ECU 10 operates the engine 1 based on thesecond map 502. By the engine 1 performing the SPCCI combustion in aportion of the operating ranges, the fuel efficiency of the engine 1improves.

When the wall temperature of the combustion chamber 17 is higher thanthe given first wall temperature and the intake air temperature ishigher than the given first the intake air temperature, the stable SPCCIcombustion of the mixture gas leaner than the stoichiometric air fuelratio can be carried out. Thus, in addition to Layer 1 and Layer 2,Layer 3 is selected, and the ECU 10 operates the engine 1 based on thefirst map 501. By the engine 1 performing the SPCCI combustion of thelean mixture gas in a portion of the operating ranges, the fuelefficiency of the engine 1 further improves.

Next, one example of control related to the layer selection of the mapexecuted by the ECU 10 will be described with reference to a flowchartof FIG. 10. First, at Step S1 after the control is started, the ECU 10reads the signals of the sensors SW1-SW17. At the following Step S2, theECU 10 determines whether the wall temperature of the combustion chamber17 is 30° C. or higher and the intake air temperature is 25° C. orhigher. If the determination at Step S2 is YES, the control shifts theprocess to Step S3, and on the other hand, if NO, the control shifts theprocess to Step S5. The ECU 10 selects only Layer 1 at Step S5. The ECU10 operates the engine 1 based on the third map 503. The control thenreturns the process.

At Step S3, the ECU 10 determines whether the wall temperature of thecombustion chamber 17 is 80° C. or higher and the intake air temperatureis 50° C. or higher. If the determination at Step S3 is YES, the controlshifts the process to Step S4, and on the other hand, if NO, the controlshifts the process to Step S6.

The ECU 10 selects Layer 1 and Layer 2 at Step S6. The ECU 10 operatesthe engine 1 based on the second map 502. The control then returns theprocess.

The ECU 10 selects Layer 1, Layer 2, and Layer 3 at Step S4. The ECU 10operates the engine 1 based on the first map. The control then returnsthe process.

(Valve Timing of Intake Valve and Exhaust Valve)

FIG. 11 illustrates one example of a change in the valve open timing IVOof the intake valve 21 when the ECU 10 controls the intake-side electricS-VT 23 according to the control logic set for Layer 2. The upper graphof FIG. 11 (i.e., a graph 1101) illustrates a change of the valve opentiming IVO of the intake valve 21 (vertical axis) versus the engine load(horizontal axis). The solid line corresponds to a case where the speedof the engine 1 is a relatively low first engine speed, and a brokenline corresponds to a case where the speed of the engine 1 is arelatively high second engine speed (first engine speed<second enginespeed).

The lower graph of FIG. 11 (i.e., a graph 1102) illustrates a change ofthe valve open timing IVO of the intake valve 21 (vertical axis) versusthe speed of the engine 1 (horizontal axis). The solid line correspondsto a case where the engine load is a relatively low first load, and thebroken line corresponds to a case where the engine load is a relativelyhigh second load (first load<second load).

In the graph 1101 and the graph 1102, the valve open timing IVO of theintake valve 21 is advanced as it goes upward and the positive overlapperiod where both the intake valve 21 and the exhaust valve 22 openbecomes longer. Therefore, the amount of the EGR gas introduced into thecombustion chamber 17 increases.

In Layer 2, the engine 1 operates with the A/F of the mixture gas at thestoichiometric air fuel ratio or the substantially stoichiometric airfuel ratio, and the G/F leaner than the stoichiometric air fuel ratio.When the engine load is low, the fuel supply amount decreases. Asillustrated in the graph 1101, when the engine load is low, the ECU 10sets the valve open timing IVO of the intake valve 21 at a timing on theretard side. Thus, the amount of the EGR gas introduced into thecombustion chamber 17 is regulated to secure the combustion stability.

Since the fuel supply amount increases when the engine load increases,the combustion stability improves. The ECU 10 sets the valve open timingof the intake valve 21 at a timing on the advance side. The pumping lossof the engine 1 can be lowered by increasing the amount of the EGR gasintroduced into the combustion chamber 17.

When the engine load further increases, the temperature inside thecombustion chamber 17 further increases. Then, the amount of theinternal EGR gas is reduced and the amount of the external EGR gas isincreased so that the temperature inside the combustion chamber 17 doesnot become too high. Therefore, the ECU 10 sets the valve open timing ofthe intake valve 21 again at a timing on the retard side.

When the engine load further increases and the supercharger 44 startsboosting, the ECU 10 sets the valve open timing of the intake valve 21again at a timing on the advance side. Since the positive overlap periodwhere both the intake valve 21 and the exhaust valve 22 open isprovided, the residual gas in the combustion chamber 17 can be purged.

Note that when the engine speed is high and low, the tendency of thechange in the valve open timing of the intake valve 21 is almost thesame.

As illustrated in the graph 1102, when the engine speed is low, the flowinside the combustion chamber 17 becomes weaker. Since the combustionstability falls, the amount of the EGR gas introduced into thecombustion chamber 17 is regulated. The ECU 10 sets the valve opentiming of the intake valve 21 at a timing on the retard side.

Since the flow inside the combustion chamber 17 becomes strong when theengine speed increases, the amount of the EGR gas introduced into thecombustion chamber 17 can be increased. The ECU 10 sets the valve opentiming of the intake valve 21 at a timing on the advance side.

When the engine speed further increases, the ECU 10 sets the valve opentiming of the intake valve 21 at a timing on the retard side accordingto the engine speed. Thus, the amount of gas introduced into thecombustion chamber 17 is maximized.

FIG. 12 illustrates one example of a change in the valve open timing IVOof the intake valve 21, the valve close timing EVC of the exhaust valve22, and an overlap period O/L of the intake valve 21 and the exhaustvalve 22, when the ECU 10 controls the intake-side electric S-VT 23 andthe exhaust-side electric S-VT 24 according to the control logic set forLayer 3.

The upper graph of FIG. 12 (i.e., a graph 1201) illustrates a change inthe valve open timing IVO of the intake valve 21 (vertical axis) versusthe engine load (horizontal axis). The solid line corresponds to a casewhere the engine speed is a relatively low third engine speed, and thebroken line corresponds to a case where the engine speed is relativelyhigh fourth engine speed (third engine speed<fourth engine speed).

The middle figure of FIG. 12 (i.e., a graph 1202) illustrates a changein the valve close timing EVC of the exhaust valve 22 (vertical axis)versus the engine load (horizontal axis). The solid line corresponds toa case where the engine speed is at the third engine speed, and thebroken line corresponds to a case where the engine speed is at thefourth engine speed.

The lower graph of FIG. 12 (i.e., a graph 1203) illustrates a change inthe overlap period O/L of the intake valve 21 and the exhaust valve 22(vertical axis) versus the engine load (horizontal axis). The solid linecorresponds to a case where the engine speed is at the third enginespeed, and the broken line corresponds to a case where the engine speedis at the fourth engine speed.

In Layer 3, the engine 1 operates by carrying out the SPCCI combustionof the mixture gas with the A/F leaner than the stoichiometric air fuelratio. When the engine load is low, the fuel supply amount decreases.Thus, the ECU 10 regulates the amount of gas introduced into thecombustion chamber 17 so that the A/F of the mixture gas does not becometoo low. As illustrated in the graph 1201, the ECU 10 sets the valveopen timing IVO of the intake valve 21 at a timing on the retard side ofan exhaust top dead center. The valve close timing of the intake valve21 becomes after an intake bottom dead center, so-called “late close”.Moreover, when the engine load is low, the ECU 10 regulates the amountof the internal EGR gas introduced into the combustion chamber 17. Asillustrated in the graph 1202, the ECU 10 sets the valve close timingEVC of the exhaust valve 22 at a timing on the advance side. The valveclose timing EVC of the exhaust valve 22 approaches an exhaust top deadcenter.

Since the fuel supply amount increases when the engine load increases,the ECU 10 does not regulate the amount of gas introduced into thecombustion chamber 17. Moreover, in order to stabilize the SPCCIcombustion of the mixture gas leaner than the stoichiometric air fuelratio, the ECU 10 increases the amount of the internal EGR gasintroduced into the combustion chamber 17. The ECU 10 sets the valveopen timing IVO of the intake valve 21 at a timing on the advance sideof an exhaust top dead center. Moreover, the ECU 10 sets the valve closetiming EVC of the exhaust valve 22 at a timing on the retard side of theexhaust top dead center. As a result, as illustrated in the graph 1203,when the engine load increases, the overlap period where both the intakevalve 21 and the exhaust valve 22 open becomes longer.

When the engine load further increases, the ECU 10 reduces the amount ofthe internal EGR gas introduced into the combustion chamber 17 so thatthe temperature inside the combustion chamber 17 does not become toohigh. The ECU 10 brings the valve close timing EVC of the exhaust valve22 closer to an exhaust top dead center. Thus, the overlap period of theintake valve 21 and the exhaust valve 22 becomes shorter. Moreover, whenthe engine load is high and the engine speed is high, the ECU 10 setsthe valve open timing of the intake valve 21 on the retard side morethan when the engine speed is low. Thus, the amount of gas introducedinto the combustion chamber 17 is maximized.

Note that in the low load area surrounded by a one-dot chain line inFIG. 12, the engine 1 may perform a reduced-cylinder operation in orderto improve fuel efficiency. When performing the reduced-cylinderoperation, the amount of gas and the internal EGR gas introduced intothe combustion chamber 17 are not regulated. As illustrated by a two-dotchain line in the graphs 1201 and 1202, the ECU 10 may set the valveopen timing of the intake valve 21 at a timing on the advance side andthe valve close timing of the exhaust valve 22 at a timing on the retardside.

(Engine Control Logic)

FIG. 13 is a flowchart illustrating the control logic of the engine 1.The ECU 10 operates the engine 1 according to the control logic storedin the memory 102. For example, the ECU 10 determines the operatingstate of the engine 1 based on the signals of the sensors SW1-SW17, andperforms calculations for adjusting properties in the combustion chamber17, the injection amount, the injection timing, and the ignition timingso that the combustion in the combustion chamber 17 becomes combustionat the SI ratio according to the operating state.

The ECU 10 first reads the signals of the sensors SW1-SW17 at Step S131.Subsequently, at Step S132, the ECU 10 determines the operating state ofthe engine 1 based on the signals of the sensors SW1-SW17, and sets atarget SI ratio (i.e., a target heat amount ratio). The target SI ratiois set according to the operating state of the engine 1.

FIG. 14 schematically illustrates one example of a setting of the targetSI ratio. When the engine load is low, the ECU 10 sets the target SIratio low, and on the other hand, when the engine load is high, it setsthe target SI ratio high. When the engine load is low, both thereduction of combustion noise and the improvement in fuel efficiency canbe achieved by increasing the ratio in the CI combustion to the SPCCIcombustion. When the engine load is high, it becomes advantageous forthe reduction of the combustion noise by increasing the ratio of the SIcombustion to the SPCCI combustion.

Returning to the flowchart of FIG. 13, the ECU 10 sets the targetin-cylinder properties for achieving the target SI ratio setting basedon the combustion model stored in the memory 102 at the following StepS133. For example, the ECU 10 sets a target temperature, a targetpressure, and target properties in the combustion chamber 17. At StepS134, the ECU 10 sets an opening of the EGR valve 54, an opening of thethrottle valve 43, an opening of the air bypass valve 48, an opening ofthe swirl control valve 56, and phase angles of the intake-side electricS-VT 23, and the exhaust-side electric S-VT 24 (i.e., a valve timing ofthe intake valve 21 and a valve timing of the exhaust valve 22), whichare required for achieving the target in-cylinder properties. The ECU 10sets the control amounts of these devices based on the map stored in thememory 102. The ECU 10 outputs signals to the EGR valve 54, the throttlevalve 43, the air bypass valve 48, the swirl control valve (SCV) 56, theintake-side electric S-VT 23, and the exhaust-side electric S-VT 24based on the control amount setting. By each device operating based onthe signal of the ECU 10, the properties in the combustion chamber 17become the target properties.

The ECU 10 further calculates predicted values or estimated values ofthe properties in the combustion chamber 17 based on the control amountsetting of each device. The predicted property value is a predictedvalue of the property in the combustion chamber 17 before the intakevalve 21 is closed. The predicted property value is used for setting ofthe fuel injection amount during the intake stroke as will be describedlater. The estimated property value is an estimated value of theproperty in the combustion chamber 17 after the intake valve 21 isclosed. The estimated property value is used for setting of the fuelinjection amount during a compression stroke, and setting of theignition timing, as will be described later. The estimated propertyvalue is also used for calculation of a property error of comparisonwith an actual combustion state.

At Step S135, the ECU 10 is sets a fuel injection amount in the intakestroke based on the predicted property value. When performing a dividedinjection during the intake stroke, the ECU 10 sets the injection amountof each injection. Note that when fuel is not injected in the intakestroke, the injection amount of fuel is zero. At Step S136, the ECU 10outputs a signal to the injector 6 so that the injector 6 injects fuelinto the combustion chamber 17 at given injection timing(s).

At Step S137, the ECU 10 is sets a fuel injection amount in acompression stroke based on the estimated property value and theinjection result of the fuel in the intake stroke. Note that when fuelis not injected in the compression stroke, the injection amount of fuelis zero. At Step S138, the ECU 10 outputs a signal to the injector 6 sothat the injector 6 injects fuel into the combustion chamber 17 at aninjection timing based on the preset map.

At Step S139, the ECU 10 sets an ignition timing based on the estimatedproperty value and the injection result of the fuel in the compressionstroke. At Step S1310, the ECU 10 outputs a signal to the ignition plug25 so that the ignition plug 25 ignites the mixture gas inside thecombustion chamber 17 at the set ignition timing.

By the ignition plug 25 igniting the mixture gas, the SI combustion orthe SPCCI combustion is performed inside the combustion chamber 17. AtStep S1311, the ECU 10 reads the change in the pressure inside thecombustion chamber 17 measured by the pressure indicating sensor SW6,and based on the change, the ECU 10 determines a combustion state of themixture gas inside the combustion chamber 17. At Step S1312, the ECU 10compares the measurement result of the combustion state with theestimated property values estimated at Step S134, and calculates anerror between the estimated properties value and the actual properties.The calculated error is used for the estimation at Step S134 in thesubsequent cycles. The ECU 10 adjusts the openings of the throttle valve43, the EGR valve 54, the swirl control valve 56, and/or the air bypassvalve 48, and the phase angles of the intake-side electric S-VT 23 andthe exhaust-side electric S-VT 24 so as to eliminate the property error.Thus, the amount of the fresh air and the EGR gas introduced into thecombustion chamber 17 are adjusted.

If the ECU 10 estimates that the temperature inside the combustionchamber 17 will be lower than the target temperature again based on theestimated property values, it advances, at Step S138, the injectiontiming in the compression stroke more than the injection timing based onthe map so that the advancing of the ignition timing becomes possible.On the other hand, if ECU 10 estimates that the temperature inside thecombustion chamber 17 will be higher than the target temperature basedon the estimated property values, it retards, at Step S138, theinjection timing in the compression stroke more than the injectiontiming based on the map so that the retarding of the ignition timingbecomes possible.

That is, if the temperature inside the combustion chamber 17 is low, aself-ignition timing θ_(CI) (see FIG. 5) of the unburnt mixture gas isdelayed after the SI combustion begins by jump-spark ignition, whichdeviates the SI ratio from the target SI ratio. In this case, anincrease in the unburnt fuel and a decrease in emission performance willbe caused.

Thus, when the ECU 10 estimates that the temperature inside thecombustion chamber 17 will be lower than the target temperature, itadvances the injection timing, and advances the ignition timing at StepS1310. Since the heat release which is sufficient for the SI combustionbecomes possible by the start of the SI combustion is made earlier, itcan prevent that the self-ignition timing θ_(CI) a of the unburntmixture gas is delayed when the temperature inside the combustionchamber 17 is low. As a result, the SI ratio approaches the target SIratio.

On the other hand, when the temperature inside the combustion chamber 17is high, the unburnt mixture gas will carry out the self ignitionshortly after SI combustion begins by the jump-spark ignition, whichdeviates the SI ratio from the target SI ratio. In this case, thecombustion noise increases.

Thus, when the ECU 10 estimates that the temperature inside thecombustion chamber 17 will be higher than the target temperature, itretards the injection timing, and retards the ignition timing at StepS1310. Since the start of SI combustion is delayed, when the temperatureinside the combustion chamber 17 is high, it can be prevented that theself-ignition timing θ_(CI) of the unburnt mixture gas becomes early. Asa result, the SI ratio approaches the target SI ratio.

The control logic of the engine 1 is configured to adjust the SI ratioby using a property setting device including the throttle valve 43, theEGR valve 54, the air bypass valve 48, the swirl control valve 56, theintake-side electric S-VT 23, and the exhaust-side electric S-VT 24. Byadjusting the properties in the combustion chamber 17, a roughadjustment of the SI ratio is possible. The control logic of the engine1 is also configured to adjust the SI ratio by adjusting the injectiontiming and the ignition timing of fuel. By adjusting the injectiontiming and the ignition timing, a difference between the cylinders canbe corrected, and a fine adjustment of the self-ignition timing can beperformed, for example. By adjusting the SI ratio by two steps, theengine 1 can achieve the target SPCCI combustion accuratelycorresponding to the operating state.

The ECU 10 also outputs signals to at least the EGR system 55 and theignition plug 25 so that the actual SI ratio by combustion becomes thetarget SI ratio. Moreover, as described above, when the engine load ishigh, since the ECU 10 makes the target SI ratio higher than when theload is low, it then outputs the signals to at least the EGR system 55and the ignition plug 25 when the engine load is high so that the SIratio becomes higher than when the load is low.

Note that the control of the engine 1 executed by the ECU 10 is notlimited to the control logic based on the combustion model describedabove.

(Method of Implementing Control Logic of Engine)

When implementing the control logic of the engine 1 for performing theSPCCI combustion described above, the parameter related to the controlamount of each device is set. For example, if the device is the ignitionplug 25, ignition energy and the ignition timing corresponding to theoperating state of the engine 1 are set. If the device is theintake-side electric S-VT 23, the valve timing of the intake valve 21corresponding to the operating state of the engine 1 is set. Below, thesetting of the valve timing of the intake valve 21 in the method ofimplementing the control logic of the engine 1 will be described withreference to the drawings.

(1) Implementing Method According to Valve Close Timing of Intake Valve

In the engine 1 for executing the SPCCI combustion, in order to reducethe combustion noise and achieve the stable SPCCI combustion, thepresent inventors found out that it was necessary to adjust thetemperature inside the combustion chamber 17 to a suitable temperatureat the start timing in the CI combustion (00: see FIG. 5). That is, whenthe temperature inside the combustion chamber 17 is low, theignitability in the CI combustion falls. When the temperature inside thecombustion chamber 17 is high, the combustion noise increases.

The temperature inside the combustion chamber 17 at the start timingθ_(CI) in the CI combustion is mainly related to the effectivecompression ratio of the engine 1. The effective compression ratio ofthe engine 1 is determined by the geometric compression ratio ε and thevalve close timing IVC of the intake valve 21. In order to put theengine for executing the SPCCI combustion into practical use, thepresent inventors newly found that a suitable IVC range existed within arange of the geometric compression ratio ε where the SPCCI combustionmay occur. The technology disclosed herein is novel in that a givenrelationship was found out between the geometric compression ratio ε andthe valve close timing IVC of the intake valve 21, where the givenrelationship is required in order to put to practical use the enginewhich performs the unique combustion mode that is the SPCCI combustion.In addition, a relational expression between the geometric compressionratio ε and the valve close timing IVC of the intake valve 21, whichwill be described later, is also novel.

Further, the technology disclosed herein is also novel in that, whenimplementing the control logic of the engine 1 for executing the SPCCIcombustion, the valve close timing IVC of the intake valve 21 is setbased on the relationship between the geometric compression ratio ε andthe valve close timing IVC of the intake valve 21.

The engine 1 switches between “Layer 2” where the SPCCI combustion isperformed with the A/F of the mixture gas being made at thestoichiometric air fuel ratio or richer than the stoichiometric air fuelratio and the G/F being made leaner than the stoichiometric air fuelratio, and “Layer 3” where the SPCCI combustion is performed with theA/F of the mixture gas being made leaner than the stoichiometric airfuel ratio. The relationship between the geometric compression ratio εand the valve close timing IVC of the intake valve 21 in Layer 2 differsfrom the relationship between the geometric compression ratio ε and thevalve close timing IVC of the intake valve 21 in Layer 3.

(1-1) Relationship Between Geometric Compression Ratio and Valve CloseTiming of Intake Valve in Layer 2

FIG. 15 illustrates the characteristic of SPCCI combustion. For example,FIG. 15 illustrates the occurring range of the SPCCI combustion againstthe EGR rate (horizontal axis) in a case where the engine 1 carries outthe SPCCI combustion of the mixture gas of which the A/F is thestoichiometric air fuel ratio and the G/F is leaner than thestoichiometric air fuel ratio, similar to Layer 2. The vertical axis inthis graph is a crank angle corresponding to a combustion center ofgravity, where the combustion center of gravity advances as it goesupward in this graph.

The occurring range of the SPCCI combustion is illustrated by a hatchedrange in this graph. The occurring range of the SPCCI combustion islocated between a line of “advancing limit” and a line of “retardinglimit.” If the combustion center of gravity advances beyond the“advancing limit” line, combustion becomes abnormal, which means thatthe SPCCI combustion does not occur. Similarly, if the combustion centerof gravity retards beyond the “retarding limit” line, self ignition doesnot occur, which means that the SPCCI combustion does not occur.

The one-dot chain line in this graph illustrates the combustion centerof gravity of combustion corresponding to MBT (Minimum advance for BestTorque). Here, the combustion center of gravity of combustioncorresponding to MBT is simply referred to as “MBT.” MBT is advanced asthe EGR rate increases.

In terms of the improvement in the fuel efficiency, it is desirable tobring the combustion center of gravity of the SPCCI combustion closer tothe MBT. As the EGR rate increases, the occurring range of the SPCCIcombustion is advanced, but the interval between the “advancing limit”line and the “retarding limit” line becomes narrower, which narrows theoccurring range of the SPCCI combustion.

When the engine load is low (when “light load”), the occurring range ofthe SPCCI combustion is on the advance side. Therefore, as illustratedby a both-ends arrow in FIG. 15, the SPCCI combustion corresponding tothe MBT can be achieved by adjusting the EGR rate within a certainwidth, and adjusting the combustion center of gravity to the advanceside or the retard side.

When the engine load increases, the amount of air introduced into thecombustion chamber 17 must be increased corresponding to the increase inthe fuel supply amount. When the EGR rate is increased corresponding tothe increase in the air amount, a large amount of the EGR gas must beintroduced into the combustion chamber 17. However, because of the limitof the supercharging capability of the supercharger 44, it is difficultto introduce both the large amount of air and the large amount of theEGR gas into the combustion chamber 17. Therefore, when the engine loadincreases, the occurring range of the SPCCI combustion becomes on theretard side. If bringing the combustion center of gravity of the SPCCIcombustion closest to the MBT when the engine load is high, the SPCCIcombustion must be performed at a point Y in FIG. 15. The point Ycorresponds to the operating state of the engine 1 with the maximum loadwhere the SPCCI combustion of the mixture gas with the A/F being at thestoichiometric air fuel ratio is possible. When the engine 1 operateswith the maximum load, it is difficult to achieve the SPCCI combustioncorresponding to the MBT by adjusting the EGR rate or the combustioncenter of gravity.

The engine 1 operates over a wide operating range from the low load tothe high load in Layer 2. In Layer 2, the state where the engine 1operates with the maximum load corresponds to the operating state at thelimit where the SPCCI combustion can occur. In Layer 2, upon determiningthe relationship between ε and IVC so that the temperature of thecombustion chamber 17 at the start timing (θ_(CI)) in the CI combustionbecomes the given temperature, it is necessary to set the relationshipbased on the temperature inside the combustion chamber 17 when theengine 1 operates with the maximum load.

In order to obtain the temperature inside the combustion chamber 17 whenthe engine 1 operates with the maximum load, the present inventorsperformed the SPCCI combustion in an actual engine 1, and usedmeasurements acquired from the engine 1. For example, the presentinventors measured various parameters when the engine 1 operates withthe maximum load in Layer 2, and estimated an actual temperature insidethe combustion chamber 17 at θ_(CI) based on the measured parameter. Thepresent inventors used an average value of a plurality of estimatedtemperatures as a reference temperature Tth1. If the temperature of thecombustion chamber 17 at θ_(CI) is the reference temperature Tth1, theSPCCI combustion can be achieved in Layer 2.

Here, as described above, θ_(CI) can be adjusted by adjusting theignition timing in the SPCCI combustion. However, while the engine 1operates with the maximum load in Layer 2, it becomes impossible toadjust θ_(CI) even if the ignition timing is adjusted when thetemperature inside the combustion chamber 17 at θ_(CI) exceeds thereference temperature Tth1. On the other hand, if the temperature insidethe combustion chamber 17 at θ_(CI) is the reference temperature Tth1 orlower, the ignition timing is adjusted (e.g., the ignition timing isadvanced) to raise the temperature inside the combustion chamber 17 atθ_(CI) to the reference temperature Tth1, thereby achieving the SPCCIcombustion.

Therefore, in order to achieve SPCCI combustion in Layer 2, therelationship between ε and IVC so that the temperature of the combustionchamber 17 at θ_(CI) does not exceed the reference temperature Tth1 isrequired.

Thus, as conceptually illustrated in FIG. 16, the present inventorsperformed an estimation of the temperature of the combustion chamber 17at θ_(CI) by using a model of the engine 1, while changing the values ofthe geometric compression ratio ε and the valve close timing IVC of theintake valve 21 in a matrix comprised of the two parameters of ε andIVC. A combination of ε and IVC where the temperature of the combustionchamber 17 becomes the reference temperature Tth1 or lower can achievethe SPCCI combustion in Layer 2. As illustrated in FIG. 16, when thegeometric compression ratio ε is high and the valve close timing IVC ofthe intake valve 21 approaches an intake bottom dead center, thetemperature of the combustion chamber 17 during CI combustion exceedsthe reference temperature Tth1.

Note that FIG. 16 illustrates the matrix where IVC is set after theintake bottom dead center. Although illustration is omitted, the presentinventors also acquired a combination of and IVC where the temperaturebecomes the reference temperature Tth1 or lower, by performing anestimation of the temperature of the combustion chamber 17 at θ_(CI),similarly for the matrix where IVC is set before the intake bottom deadcenter.

A graph 1701 of the upper graph in FIG. 17 illustrates approximations(approx.) (I) and (II) calculated based on the combination of ε and IVC.The horizontal axis of the graph 1701 is the geometric compression ratioε, and the vertical axis is the valve close timing IVC (deg.aBDC) of theintake valve 21. Although illustration is omitted, the present inventorsplotted on the graph 1701 the combination of ε and IVC where thetemperature becomes the reference temperature Tth1 or lower, anddetermined the approximations (I) and (II) based on the plots.

The graph 1701 corresponds to a case where the engine speed is 2,000rpm. The approximations (I) and (II) are as follows.IVC=−0.4288ε²+31.518ε−379.88  Approximation (I):IVC=1.9163ε²−89.935ε+974.94  Approximation (II):

In the graph 1701, the combination of ε and IVC on the left side of theapproximations (I) and (II) and ε=17, the temperature of the combustionchamber 17 during the CI combustion becomes the reference temperatureTth1 or lower. In this combination, it is possible to carry out theSPCCI combustion of the mixture gas with the A/F being thestoichiometric air fuel ratio and the G/F being leaner than thestoichiometric air fuel ratio.

The relationship between ε and IVC described above is a relationshipbased on the maximum temperature of the combustion chamber 17 in Layer2.

On the other hand, in Layer 2, also while the engine 1 operates with thelight load, the relationship between ε and IVC must be defined so thatthe temperature of the combustion chamber 17 becomes the giventemperature.

The temperature of the combustion chamber 17 when the SPCCI combustionis performed is a result of two pressure buildups of a pressure buildupby the compression work of the piston 3 in a compression stroke, and thepressure buildup caused by the heat generation of the SI combustion. Thecompression work of the piston 3 is determined by the effectivecompression ratio. If the effective compression ratio is too low, thepressure buildup by the compression work of the piston 3 decreases.Unless the pressure buildup, which is caused by the heat generation ofthe SI combustion after the flame propagation in the SPCCI combustionprogresses, increases considerably, the in-cylinder temperature cannotbe raised to an ignition temperature. As a result, since the amount ofthe mixture gas which is ignited by the compressed self ignition islittle, and most of the mixture gas burn by flame propagation, thecombustion period becomes longer and fuel efficiency decreases. In orderto stabilize the CI combustion in the SPCCI combustion and maximize fuelefficiency, it is necessary to maintain the effective compression ratioabove a certain value. Therefore, the relationship between ε and IVCmust be determined accordingly.

Similarly, the present inventors measured various parameters when theactual engine 1 operates with the light load, and estimated an actualtemperature inside the combustion chamber 17 at θ_(CI) based on themeasured parameters. The present inventors used an average value of aplurality of estimated temperatures as a reference temperature Tth2.

If the temperature inside the combustion chamber 17 at θ_(CI) is thereference temperature Tth2 or higher while the engine 1 operates withthe light load, the SPCCI combustion can be achieved by delaying theignition timing. However, since the temperature of the combustionchamber 17 is too low if the temperature at θ_(CI) is lower than thereference temperature Tth2, the SPCCI combustion cannot be achieved evenif the ignition timing is advanced.

Thus, in order to achieve the SPCCI combustion in Layer 2, arelationship between and IVC so that the temperature of the combustionchamber 17 at θ_(CI) becomes the reference temperature Tth2 or higher isrequired.

The present inventors performed an estimation of the temperature of thecombustion chamber 17 during CI combustion by using a model of theengine 1, while changing the values of the geometric compression ratio εand the valve close timing IVC of the intake valve 21 in a matrixcomprised of two parameters of ε and IVC, similarly to the matrixillustrated in FIG. 16. In this matrix, the combination of ε and IVCwhere the temperature of the combustion chamber 17 becomes the referencetemperature Tth2 or higher can achieve the SPCCI combustion in Layer 2.

In the graph 1701 of FIG. 17, approximations (III) and (IV) calculatedbased on the combination of ε and IVC where the temperature becomes thereference temperature Tth2 or higher are also illustrated. Theapproximations (III) and (IV) are as follows.IVC=−0.4234ε²+22.926ε−167.84  Approximation (III):IVC=0.4234ε²−22.926ε+207.84  Approximation (IV):

In the graph 1701, the combination of ε and IVC on the right side of theapproximations (III) and (IV), the temperature of the combustion chamber17 during CI combustion becomes the reference temperature Tth2 orhigher. In this combination, the SPCCI combustion of the mixture gaswith the A/F being the stoichiometric air fuel ratio and the G/F beingleaner than the stoichiometric air fuel ratio is possible.

As seen in FIG. 17, the relationship between ε and IVC is substantiallyvertically symmetrical with respect to IVC=about 20 deg.aBDC. IVC=20deg.aBDC corresponds to a valve close timing (i.e., the best IVC) atwhich the amount of gas introduced into the combustion chamber 17becomes the maximum when the engine speed is 2,000 rpm. Moreover,IVC=120 deg.aBDC is a retarding limit of the valve close timing IVC ofthe intake valve 21, and IVC=−80 deg.aBDC is an advancing limit of thevalve close timing IVC of the intake valve 21.

The combination of ε and IVC within a range surrounded by theapproximations (I), (II), (III), and (IV) in FIG. 17 is a combinationwhich can put to practical use the engine 1 for executing the SPCCIcombustion in Layer 2. In other words, the combination of ε and IVCoutside this range cannot put the engine 1 to practical use forexecuting the SPCCI combustion in Layer 2.

The design engineer has to determine IVC within the ε-IVC valid rangehatched in FIG. 17, upon determining the valve close timing IVC of theintake valve 21 when the engine 1 operates in Layer 2.

For example, if the geometric compression ratio ε is set as 10≤ε<17, thedesign engineer determines the valve close timing IVC (deg.aBDC) so thatthe following expression is satisfied.0.4234ε²−22.926ε+207.84≤IVC≤−0.4234ε²+22.926ε−167.84  (1)

Moreover, if the geometric compression ratio ε is set as 17≤ε<20, thedesign engineer determines the valve close timing IVC (deg.aBDC) so thatthe following expression is satisfied.−0.4288ε²+31.518ε−379.88≤IVC≤−0.4234ε²+22.926ε−167.84  (2)or0.4234ε²−22.926ε+207.84≤IVC≤1.9163ε²−89.935ε+974.94  (3)

Further, if the geometric compression ratio ε is set as 20≤ε<30, thedesign engineer determines the valve close timing IVC (deg.aBDC) so thatthe following expression is satisfied.−0.4288ε²+31.518ε−379.88≤IVC≤120  (4)or−80≤IVC≤1.9163ε²−89.935ε+974.94  (5)

By setting the valve close timing IVC of the intake valve 21 based onthe relational expressions (1) to (5), the SPCCI combustion of themixture gas with the A/F being the stoichiometric air fuel ratio orricher than the stoichiometric air fuel ratio and the G/F being leanerthan the stoichiometric air fuel ratio is achieved. Note that the valveclose timing IVC is set for each operating state which is determined bythe load and the engine speed in Layer 2. The example illustrated by thesolid line in FIG. 17 is the ε-IVC valid range when the engine speed is2,000 rpm, as described above. If the engine speed changes, the ε-IVCvalid range also changes. As the engine speed increases, the best IVC isretarded.

For example, when the engine speed is 3,000 rpm, the best IVC is about22 deg.aBDC. As illustrated by broken lines in FIG. 17, the ε-IVC validrange when the engine speed is 3,000 rpm can be obtained by parallellytranslating the ε-IVC valid range when the engine speed is 2,000 rpm tothe retard side by about 2 deg.

Therefore, if the geometric compression ratio ε is set as 10≤ε<17, whenthe engine speed is 3,000 rpm, the design engineer determines the valveclose timing IVC (deg.aBDC) so that the following expression issatisfied.0.4234ε²−22.926ε+209.84≤IVC≤−0.4234ε²+22.926ε−165.84  (1⁻¹)

Moreover, if the geometric compression ratio ε is set as 17≤ε<20, whenthe engine speed is 3,000 rpm, the design engineer determines the valveclose timing IVC (deg.aBDC) so that the following expression issatisfied.−0.4288ε²+31.518ε−377.88≤IVC≤−0.4234ε²+22.926ε−165.84  (2⁻¹)or0.4234ε²−22.926ε+209.84≤IVC≤1.9163ε²−89.935ε+976.94  (3⁻¹)

Further, if the geometric compression ratio ε is set as 20≤ε≤30, thedesign engineer determines the valve close timing IVC (deg.aBDC) so thatthe following expression is satisfied.−0.4288ε²+31.518ε−377.88≤IVC≤120  (4⁻¹)or−80≤IVC≤1.9163ε²−87.935ε+976.94  (5⁻¹)

Moreover, when the engine speed is 4,000 rpm, the best IVC is about 28deg.aBDC. As illustrated by one-dot chain lines in FIG. 17, the ε-IVCvalid range when the engine speed is 4,000 rpm is obtained by parallellytranslating the ε-IVC valid range when the engine speed is 2,000 rpm tothe retard side by about 8 deg.

Therefore, if the geometric compression ratio ε is set as 10≤ε<17, whenthe engine speed is 4,000 rpm, the design engineer determines the valveclose timing IVC (deg.aBDC) so that the following expression issatisfied.0.4234ε²−22.926ε+215.84≤IVC≤−0.4234ε²+22.926ε−159.84  (1⁻²)

Moreover, if the geometric compression ratio ε is set as 17≤ε<20, whenthe engine speed is 4,000 rpm, the design engineer determines the valveclose timing IVC (deg.aBDC) so that the following expression issatisfied.−0.4288ε²+31.518ε−371.88≤IVC≤−0.4234ε²+22.926ε−159.84  (2⁻²)or0.4234ε²−22.926ε+215.84≤IVC≤1.9163ε²−89.935ε+982.94  (3⁻²)

Further, if the geometric compression ratio ε is set as 20≤ε≤30, thedesign engineer determines the valve close timing IVC (deg.aBDC) so thatthe following expression is satisfied.−0.4288ε²+31.518ε−371.88≤IVC≤120  (4⁻²)or−80≤IVC≤1.9163ε²−89.935ε+982.94  (5⁻²)

If a correction term C according to the engine speed NE (rpm) of theengine 1 is set as the following,C=3.3×10⁻¹⁰ NE ³−1.0×10⁻⁶ NE ³+7.0×10⁻⁴ NEthe relational expression of ε and IVC in Layer 2 can be expressed asfollows. If the geometric compression ratio ε is 10≤ε<17,0.4234ε²−22.926ε+207.84+C≤IVC≤−0.4234ε²+22.926ε−167.84+C  (1⁻³)If the geometric compression ratio ε is 17≤ε<20,0.4288ε²+31.518ε−379.88+C≤IVC≤−0.4234ε²+22.926ε−167.84+C  (2⁻³)or0.4234ε²−22.926ε+207.84+C≤IVC≤1.9163ε²−89.935ε+974.94+C  (3⁻³)If the geometric compression ratio ε is 20≤ε≤30,−0.4288ε²+31.518ε−379.88+C≤IVC≤120  (4⁻³)or−80≤IVC≤1.9163ε²−89.935ε+974.94+C  (5⁻³)

The design engineer determines the valve close timing IVC based on theε-IVC valid range determined for every engine speed of the engine 1. Asa result, the design engineer can set the valve timing of the intakevalve 21 in Layer 2 as illustrated in FIG. 11.

(1-2) Change in ε-IVC Valid Range by Difference of Octane Number

The graph 1701 in FIG. 17 is a relationship between ε and IVC when thefuel is the high octane fuel (octane number is about 96). A graph 1702illustrated in the lower graph is a relationship between ε and IVC whenthe fuel is the low octane fuel (octane number is about 91). Accordingto the examination of the present inventors, when the fuel was the lowoctane fuel, it was found that the ε-IVC valid range shifts by 1.3compression ratios toward the lower compression ratio from the ε-IVCvalid range of the high octane fuel.

Accordingly, upon determining the valve close timing IVC in the engine 1of the low octane fuel, if the geometric compression ratio ε is set as10≤ε<15.7, when the engine speed is 2,000 rpm, the design engineerdetermines the valve close timing IVC (deg.aBDC) so that the followingexpression is satisfied.0.4234ε²−21.826ε+178.75≤IVC≤−0.4234ε²+21.826ε−138.75  (6)

Moreover, if the geometric compression ratio ε is set as 15.7≤ε<18.7 inthe engine 1 of the low octane fuel, when the engine speed is 2,000 rpm,the design engineer determines the valve close timing IVC (deg.aBDC) sothat the following expression is satisfied.−0.5603ε²+34.859ε−377.22≤IVC≤−0.4234ε²+21.826ε−138.75  (7or0.4234ε²−21.826ε+178.75≤IVC≤1.9211ε²−85.076ε+862.01  (8

Further, if the geometric compression ratio ε is set as 18.7≤ε≤30 in theengine 1 of the low octane fuel, when the engine speed is 2,000 rpm, thedesign engineer determines the valve close timing IVC (deg.aBDC) so thatthe following expression is satisfied.−0.5603ε²+34.859ε−377.22≤IVC≤120  (9)or−80≤IVC≤1.9211ε²−85.076ε+862.01  (10)

The cross-hatched range in the graph 1702 of FIG. 17 is an overlappingrange of the ε-IVC valid range of the high octane fuel and the ε-IVCvalid range of the low octane fuel. The design engineer can set thecontrol logic which suits both the engine 1 using the high octane fueland the engine 1 using the low octane number fuel, if IVC is determinedwithin the overlapping range of the two valid ranges. Even if the octanenumber of fuel differs for every destination of this product, the designengineer can collectively implement the engine control logic. Thispackage design has an advantage which lessens the design time and labor.

Note that although illustration is omitted, the ε-IVC valid range isparallelly translated to the retard side also in the engine 1 of the lowoctane number fuel, when the engine speed increases. If the geometriccompression ratio ε is set as 10≤ε<15.7 in the engine 1 of the lowoctane number fuel, when the engine speed is 3,000 rpm, the designengineer determines the valve close timing IVC (deg.aBDC) so that thefollowing expression is satisfied.0.4234ε²−21.826ε+180.75≤IVC≤−0.4234ε²+21.826ε−136.75  (6⁻¹)

Moreover, if the geometric compression ratio ε is set as 15.7≤ε<18.7 inthe engine 1 of the low octane fuel, when the engine speed is 3,000 rpm,the design engineer determines the valve close timing IVC (deg.aBDC) sothat the following expression is satisfied.−0.5603ε²+34.859ε−375.22≤IVC≤−0.4234ε²+21.826ε−136.75  (7⁻¹)or0.4234ε²−21.826ε+180.75≤IVC≤1.9211ε²−85.076ε+864.01  (8⁻¹)

Further, if the geometric compression ratio ε is set as 18.7≤ε≤30 in theengine 1 of the low octane fuel, when the engine speed is 3,000 rpm, thedesign engineer determines the valve close timing IVC (deg.aBDC) so thatthe following expression is satisfied.−0.5603ε²+34.859ε−375.22≤IVC≤120  (9⁻¹)or−80≤IVC≤1.9211ε²−85.076ε+864.01  (10⁻¹)

Moreover, if the geometric compression ratio ε is set as 10≤ε<15.7 inthe engine 1 of the low octane fuel, when the engine speed is 4,000 rpm,the design engineer determines the valve close timing IVC (deg.aBDC) sothat the following expression is satisfied.0.4234ε²−21.826ε+186.75≤IVC≤−0.4234ε²+21.826ε−130.75  (6⁻²)

Moreover, if the geometric compression ratio ε is set as 15.7≤ε<18.7 inthe engine 1 of the low octane fuel, when the engine speed is 4,000 rpm,the design engineer determines the valve close timing IVC (deg.aBDC) sothat the following expression is satisfied.−0.5603ε²+34.859ε−369.22≤IVC≤−0.4234ε²+21.826ε−130.75  (7⁻²)or0.4234ε²−21.826ε+186.75≤IVC≤1.9211ε²−85.076ε+870.01  (8⁻²)

Further, if the geometric compression ratio ε is set as 18.7≤ε≤30 in theengine 1 of the low octane fuel, when the engine speed is 4,000 rpm, thedesign engineer determines the valve close timing IVC (deg.aBDC) so thatthe following expression is satisfied.−0.5603ε²+34.859ε−369.22≤IVC≤120  (9⁻²)or−80≤IVC≤1.9211ε²−77.076ε+870.01  (10⁻²)

If the correction term C according to the engine speed NE (rpm) of theengine 1 is used similar to the above, the relational expression of εand IVC in Layer 2 in the engine 1 of the low octane fuel can beexpressed as follows.

If the geometric compression ratio ε is 10≤ε<15.7,0.4234ε²−21.826ε+178.75+C≤IVC≤−0.4234ε²+21.826ε−138.75+C  (6⁻³)If the geometric compression ratio ε is 15.7≤ε<18.7,−0.5603ε²+34.859ε−377.22+C≤IVC≤−0.4234ε²+21.826ε−138.75+C  (7⁻³)or0.4234ε²−21.826ε+178.75+C≤IVC≤1.9211ε²−85.076ε+862.01+C  (8⁻³)If the geometric compression ratio is 18.7≤ε≤30,−0.5603ε²+34.859ε−377.22+C≤IVC≤120  (9⁻³)or−80≤IVC≤1.9211ε²−85.076ε+862.01+C  (10⁻³)(1-3) Relationship Between Geometric Compression Ratio and Valve CloseTiming of Intake Valve in Layer 3

FIG. 18 illustrates the characteristic of the SPCCI combustion when theengine 1 carries out the SPCCI combustion of the mixture gas where theA/F is leaner than the stoichiometric air fuel ratio, similar to Layer3. FIG. 18 illustrates a range where the SPCCI combustion is stable withrespect to the G/F (horizontal axis). The vertical axis in this graph isa crank angle corresponding to a combustion center of gravity, and thecombustion center of gravity is advanced as it goes upward in thisgraph.

The range where the SPCCI combustion is stabilized is a range surroundedby a curve in this graph. When the engine load is low, the range wherethe SPCCI combustion is stabilized is located at upper left in FIG. 18.When the engine load increases, the range where the SPCCI combustion isstabilized moves downwardly in FIG. 18.

FIG. 18 also illustrates a range where discharge of raw NO_(x) can bereduced. The range where the discharge of raw NO_(x) can be reduced islocated at lower right in FIG. 18. This range has a triangular shape inFIG. 18. If the A/F is leaner than the stoichiometric air fuel ratio,raw NO_(x) cannot be purified by the three-way catalyst. In Layer 3, theengine 1 must satisfy both securing of the stability of the SPCCIcombustion and reduction of the discharge of raw NO_(x).

As seen in this graph, if the engine load is high, an overlapping areaof the range where the combustion stability is secured and the rangewhere the discharge of raw NO_(x) can be reduced increases. On the otherhand, if the engine load is low, the overlapping area of the range wherethe combustion stability is secured and the range where the discharge ofraw NO_(x) can be reduced decreases.

Regarding Layer 3, the state where the engine 1 operates with the lightload corresponds to a limit of the operating state where the SPCCIcombustion can occur. Regarding Layer 3, upon determining therelationship between ε and IVC so that the temperature of the combustionchamber 17 at the start timing (θ_(CI)) of the CI combustion becomes thegiven temperature, it is necessary to be set based on the temperatureinside the combustion chamber 17 when the engine 1 operates with thelight load.

Similar to above, using the actual engine 1, the present inventorsmeasured various parameters when operating the engine 1 with the lightload in Layer 3, and estimated an actual temperature inside thecombustion chamber 17 at θ_(CI) based on the measured parameters. Then,an average value of a plurality of estimated temperatures was determinedas a reference temperature Tth3. If the temperature of the combustionchamber 17 at θ_(CI) is the reference temperature Tth3, the SPCCIcombustion can be achieved in Layer 3. This reference temperature Tth3corresponds to the minimum temperature. If the temperature inside thecombustion chamber 17 at θ_(CI) is the reference temperature Tth3 orhigher while the engine 1 operates with the light load, the SPCCIcombustion can be achieved by delaying the ignition timing. However, ifthe temperature inside the combustion chamber 17 at θ_(CI) is lower thanthe reference temperature Tth3, the SPCCI combustion cannot be achievedeven if the ignition timing is advanced.

Therefore, in order to achieve the stable SPCCI combustion in Layer 3,the relationship between ε and IVC so that the temperature inside thecombustion chamber 17 at θ_(CI) becomes the reference temperature Tth3or higher is required.

Thus, as conceptually illustrated in FIG. 19, the present inventorsestimated the temperature of the combustion chamber 17 during the CIcombustion by using the model of the engine 1, while changing the valuesof IVC and O/L in a matrix of the valve close timing IVC of the intakevalve 21 and the overlap period O/L of the intake valve 21 and theexhaust valve 22, for every geometric compression ratio ε (ε1, ε2 . . .) (reference numeral 1901). As hatched in the matrix of the referencenumeral 1901, a combination of IVC and O/L where the temperature becomesthe reference temperature Tth3 or higher can achieve the SPCCIcombustion in Layer 3.

Moreover, in order to reduce the discharge of raw NO_(x), the G/F of themixture gas must be made a given value or higher. As illustrated by areference numeral 1902 in FIG. 19, the present inventors estimated theG/F by using the model of the engine 1, while changing the values of thevalve close timing IVC and the overlap period O/L in the matrixcomprised of the two parameters of IVC and O/L. As oblique lines aredrawn in the matrix of the reference numeral 1902, the combination ofIVC and O/L where the G/F becomes the given value or higher can reducethe discharge of raw NO_(x).

Then, the present inventors determined the relationship between ε andIVC which can achieve both the stability of the SPCCI combustion and thereduction of raw NO_(x) discharge, by overlapping the combination of IVCand O/L where the temperature becomes the reference temperature Tth3 orhigher which is illustrated by the reference numeral 1901, and thecombination of IVC and O/L where the G/F becomes the given value orhigher which is illustrated by the reference numeral 1902. That is, inthe matrix of a reference numeral 1903, a cross-hatched range is thecombination of ε and IVC which can achieve both the stability of theSPCCI combustion and the reduction of NO_(x) discharge.

Note that although illustration is omitted, the present inventorsacquired a combination of IVC and O/L where the temperature becomes thereference temperature Tth3 or higher and a combination of IVC and O/Lwhere the G/F becomes the given value or higher by estimating thetemperature of the combustion chamber 17 and the G/F during the CIcombustion, similarly for the matrix where the valve close timing of theintake valve 21 is set before the intake bottom dead center.

FIG. 20 illustrates approximations (V) and (VI) calculated based on thecombinations of ε and IVC. The horizontal axis in FIG. 20 is thegeometric compression ratio ε, and the vertical axis is the valve closetiming IVC (deg. aBDC) of the intake valve 21.

The upper graph 2001 of FIG. 20 corresponds to a case when the enginespeed is 2,000 rpm. The approximations (V) and (VI) are as follows.IVC=0.9949ε²+41.736ε−361.16  Approximation (V):IVC=0.9949ε²−41.736ε+401.16  Approximation (VI):

In FIG. 20, the combinations of ε and IVC on the right side of theapproximations (V) and (VI) have the temperature of the combustionchamber 17 during CI combustion becoming the reference temperature Tth3or higher, thereby achieving the SPCCI combustion of the mixture gaswith the A/F being leaner than the stoichiometric air fuel ratio.

In Layer 3, the relationship between ε and IVC described above is arelationship based on the minimum temperature of the combustion chamber17 which can achieve the SPCCI combustion when the engine 1 operateswith the light load.

On the other hand, if the temperature inside the combustion chamber 17is too high, the CI combustion begins before the start of the SIcombustion and the SPCCI combustion cannot be performed, regardless ofLayer 2 or Layer 3.

Here, the concept of the SPCCI combustion is such that, as describedabove, when the ignition plug 25 ignites the mixture gas, the mixturegas around the ignition plug 25 starts the SI combustion, and, afterthat, surrounding mixture gas carries out the CI combustion. From theexaminations by experiments, etc. which the present inventors conducteduntil now, it was found that the self ignition of the mixture gasoccurred when the surrounding temperature of the mixture gas whichself-ignites exceeds a given reference temperature Tth4, and thisreference temperature Tth4 was not necessarily in agreement with a meantemperature of the entire combustion chamber 17. From this knowledge, ifthe mean temperature inside the combustion chamber 17 at a compressiontop dead center reaches the reference temperature Tth4, it is thoughtthat the CI combustion will begin before the SI combustion begins, andin this case, the SPCCI combustion cannot be performed.

Thus, the present inventors estimated the temperature of the combustionchamber 17 at the compression top dead center by using the model of theengine 1, while changing the values of the valve close timing IVC andthe overlap period O/L in the matrix of the valve close timing IVC ofthe intake valve 21, and the overlap period O/L of the intake valve 21and the exhaust valve 22, for every geometric compression ratio ε (ε1,ε2 . . . ), similar to the matrix of the reference numeral 1901 of FIG.19. The combination of IVC and O/L where the temperature inside thecombustion chamber 17 at the compression top dead center exceeds thereference temperature Tth4 cannot achieve the SPCCI combustion, but thecombination of IVC and O/L being the reference temperature Tth4 or lowercan achieve the SPCCI combustion.

FIG. 20 illustrates approximations (VII) and (VIII) calculated based onthe combination of ε and IVC where the temperature inside the combustionchamber 17 at the compression top dead center becomes the referencetemperature Tth4 or lower. The approximations (VII) and (VIII) are asfollows.IVC=−4.7481ε²+266.75ε−3671.2  Approximation (VII):andIVC=4.7481ε²−266.75ε+3711.2  Approximation (VIII):

In FIG. 20, the combination of ε and IVC on the left side theapproximations (VII) and (VIII) can avoid that the CI combustion beginsbefore the SI combustion, and achieves the SPCCI combustion.

As seen in FIG. 20, also in Layer 3, the relationship between ε and IVCis substantially vertically symmetrical with respect to IVC=20 deg.aBDC.Moreover, IVC=75 deg.aBDC is a retarding limit of the valve close timingof the intake valve 21 set in consideration of the amount of gasintroduced into the combustion chamber 17 when the engine 1 operates inLayer 3. Similarly, IVC=−35 deg.aBDC is an advancing limit of the valveclose timing of the intake valve 21 set in consideration of the amountof gas introduced into the combustion chamber 17.

When determining the valve close timing IVC of the intake valve 21 inthe case where the engine 1 operates in Layer 3, the design engineermust determine IVC within the ε-IVC valid range surrounded by theapproximations (V), (VI), (VII), and (VIII) in FIG. 20 (the hatchedrange in FIG. 20).

For example, if the geometric compression ratio ε is set as 10≤ε<20,when the engine speed is 2,000 rpm, the design engineer determines thevalve close timing IVC (deg.aBDC) so that the following expression issatisfied.0.9949ε²−41.736ε+401.16≤IVC≤−0.9949ε²+41.736ε−361.16  (11)

Moreover, if the geometric compression ratio ε is set as 20≤ε<25, whenthe engine speed is 2,000 rpm, the design engineer determines the valveclose timing IVC (deg.aBDC) so that the following expression issatisfied.−35≤IVC≤75  (12)

Further, if the geometric compression ratio ε is set as 25≤ε≤30, whenthe engine speed is 2,000 rpm, the design engineer determines the valveclose timing IVC (deg.aBDC) so that the following expression issatisfied.−4.7481ε²+266.75ε−3671.2≤IVC≤75  (13)or−35≤IVC≤4.7481ε²−266.75ε+3711.2  (14)

By setting the valve close timing IVC of the intake valve 21 based onthe relational expressions (11) to (14), the SPCCI combustion of themixture gas with that the A/F being leaner than the stoichiometric airfuel ratio carries is achieved. Note that the valve close timing IVC isset for each operating state which is determined by the load and theengine speed in Layer 3.

The example illustrated by a solid line in FIG. 20 is the ε-IVC validrange when the engine speed is 2,000 rpm, as described above. If theengine speed changes, the ε-IVC valid range also changes. When theengine speed increases, the ε-IVC valid range is also parallellytranslated to the retard side, similarly in FIG. 20. Therefore, if thegeometric compression ratio is set as 10≤ε<20, when the engine speed is3,000 rpm (see a broken line), the design engineer determines the valveclose timing IVC (deg.aBDC) so that the following expression issatisfied.0.9949ε²−41.736ε+403.16≤IVC≤−0.9949ε²+41.736ε−359.16  (11⁻¹)

If the geometric compression ratio ε is set as 20≤ε<25, when the enginespeed is 3,000 rpm, the design engineer determines the valve closetiming IVC (deg.aBDC) so that the following expression is satisfied.−33≤IVC≤77  (12⁻¹)

Further, if the geometric compression ratio ε is set as 25≤ε≤30, whenthe engine speed is 3,000 rpm, the design engineer determines the valveclose timing IVC (deg.aBDC) so that the following expression issatisfied.−4.7481ε²+266.75ε−3669.2≤IVC≤77  (13⁻¹)or−33≤IVC≤4.7481ε²−266.75ε+3713.2  (14⁻¹)

If the geometric compression ratio ε is set as 10≤ε<20, when the enginespeed is 4,000 rpm (see one-dot chain line), the design engineerdetermines the valve close timing IVC (deg.aBDC) so that the followingexpression is satisfied.0.9949ε²−41.736ε+409.16≤IVC≤−0.9949ε²+41.736ε−353.16  (11)

If the geometric compression ratio ε is set as 20≤ε<25, when the enginespeed is 4,000 rpm, the design engineer determines the valve closetiming IVC (deg.aBDC) so that the following expression is satisfied.−27≤IVC≤83  (12⁻²)

Further, if the geometric compression ratio ε is set as 25≤ε≤30, whenthe engine speed is 4,000 rpm, the design engineer determines the valveclose timing IVC (deg.aBDC) so that the following expression issatisfied.−4.7481ε²+266.75ε−3663.2≤IVC≤83  (13⁻²)or−27≤IVC≤4.7481ε²−266.75ε+3719.2  (14⁻²)

If the correction term C according to the engine speed NE (rpm) of theengine 1 is used similar to the above, the relational expression of εand IVC in Layer 3 can be expressed as follows. If the geometriccompression ratio ε is 10≤ε<20,0.9949ε²−41.736ε+401.16+C≤IVC≤−0.9949ε²+41.736ε−361.16+C  (11⁻³)If the geometric compression ratio ε is 20≤ε<25,−35+C≤IVC≤75+C  (12⁻³)If the geometric compression ratio ε is 25≤ε≤30,−4.7481ε²+266.75ε−3671.2+C≤IVC≤75+C  (13⁻³)or−35+C≤IVC≤4.7481ε²−266.75ε+3711.2+C  (14⁻³)

The design engineer determines the valve close timing IVC based on theε-IVC valid range determined for every engine speed of the engine 1. Asa result, the design engineer can set the valve timing of the intakevalve 21 in Layer 3 as illustrated in FIG. 12.

Moreover, the lower graph 2002 of FIG. 20 is a relationship between εand IVC when the fuel is the low octane fuel.

Upon determining the valve close timing IVC in the engine 1 of the lowoctane fuel, if the geometric compression ratio ε is set as 10≤ε<18.7,when the engine speed is 2,000 rpm, the design engineer determines thevalve close timing IVC (deg.aBDC) so that the following expression issatisfied.0.9949ε²−39.149ε+348.59≤IVC≤−0.9949ε²+39.149ε−308.59  (15)

Moreover, if the geometric compression ratio ε is set as 18.7≤ε<23.7 inthe engine 1 of the low octane fuel, when the engine speed is 2,000 rpm,the design engineer determines the valve close timing IVC (deg.aBDC) sothat the following expression is satisfied.−35≤IVC≤75  (16)

Further, if the geometric compression ratio ε is set as 23.7≤ε≤30 in theengine 1 of the low octane fuel, when the engine speed is 2,000 rpm, thedesign engineer determines the valve close timing IVC (deg.aBDC) so thatthe following expression is satisfied.−3.1298ε²+172.48ε−2300≤IVC≤75  (17)or−35≤IVC≤3.1298ε²−172.48ε+2340  (18)

The cross-hatched range in the lower graph 2002 of FIG. 20 is anoverlapping range of the ε-IVC valid range of the high octane fuel andthe ε-IVC valid range of the low octane fuel. Similar to the above, thedesign engineer can set the control logic which suits both the engine 1using the high octane fuel and the engine 1 using the low octane fuel bydetermining IVC within the overlapping range of the two occurringranges.

Note that although illustration is omitted, if the engine speedincreases, the ε-IVC valid range is parallelly translated to the retardside also in the engine 1 of the low octane fuel. If the geometriccompression ratio ε is set as 10<ε<18.7 in the engine 1 of the lowoctane fuel, when the engine speed is 3,000 rpm, the design engineerdetermines the valve close timing IVC (deg.aBDC) so that the followingexpression is satisfied.0.9949ε²−39.149ε+350.59≤IVC≤−0.9949ε²+39.149ε−306.59  (15⁻¹)

Moreover, if the geometric compression ratio ε is set as 18.7≤ε<23.7 inthe engine 1 of the low octane fuel, when the engine speed is 3,000 rpm,the design engineer determines the valve close timing IVC (deg.aBDC) sothat the following expression is satisfied.−33≤IVC≤77  (16⁻¹)

Further, if the geometric compression ratio ε is set as 23.7≤ε≤30 in theengine 1 of the low octane fuel, when the engine speed is 3,000 rpm, thedesign engineer determines the valve close timing IVC (deg.aBDC) so thatthe following expression is satisfied.−3.1298ε²+172.48ε−2298≤IVC≤77  (17⁻¹)or−33≤IVC≤3.1298ε²−172.48ε+2342  (18⁻¹)

If the geometric compression ratio ε is set as 10≤ε<18.7 in the engine 1of the low octane fuel, when the engine speed is 4,000 rpm, the designengineer determines the valve close timing IVC (deg.aBDC) so that thefollowing expression is satisfied.0.9949ε²−39.149ε+356.59≤IVC≤−0.9949ε²+39.149ε−300.59  (15⁻²)

Moreover, if the geometric compression ratio ε is set as 18.7≤ε<23.7 inthe engine 1 of the low octane fuel, when the engine speed is 4,000 rpm,the design engineer determines the valve close timing IVC (deg.aBDC) sothat the following expression is satisfied.−27≤IVC≤83  (16⁻²)

Further, if the geometric compression ratio ε is set as 23.7≤ε≤30 in theengine 1 of the low octane fuel, when the engine speed is 4,000 rpm, thedesign engineer determines the valve close timing IVC (deg.aBDC) so thatthe following expression is satisfied.−3.1298ε²+172.48ε−2292≤IVC≤83  (17⁻²)or−27≤IVC≤3.1298ε²−172.48ε+2348  (18⁻²)

If the correction term C according to the engine speed NE (rpm) of theengine 1 is used similar to the above, the relational expression of εand IVC in Layer 3, in the engine 1 of the low octane fuel can beexpressed as follows. If the geometric compression ratio ε is 10≤ε<18.7,0.9949ε²−39.149ε+348.59+C≤IVC≤−0.9949ε²+39.149ε−308.59+C  (15⁻³)If the geometric compression ratio ε is 18.7≤ε<23.7,−35+C≤IVC≤75+C  (16⁻³)If the geometric compression ratio ε is 23.7≤ε≤30,−3.1298ε²+172.48ε−2300+C≤IVC≤75+C  (17⁻³)or−35+C≤IVC≤3.1298ε²−172.48ε+2340+C  (18⁻³)(1-4) Relationship Between Geometric Compression Ratio and Valve CloseTiming of Intake Valve in Layers 2 and 3

FIG. 21 illustrates a relationship between the geometric compressionratio ε and the valve close timing IVC of the intake valve 21 where theSPCCI combustion is possible in both Layer 2 and Layer 3. Thisrelational expression is obtained from the ε-IVC valid range of FIG. 17and the ε-IVC valid range of FIG. 20.

When the ECU 10 selects Layer 3 according to the temperature, etc. ofthe engine 1, the low-load operating range of the engine 1 is switchedfrom Layer 2 to Layer 3. If the valve close timing IVC of the intakevalve 21 is set so that the SPCCI combustion is possible in both Layer 2and Layer 3, it is possible to continuously perform the SPCCI combustioneven when the map of the engine 1 is switched from Layer 2 to Layer 3.

An upper graph 2101 of FIG. 21 is a relationship between ε and IVC whenthe fuel is the high octane fuel. A lower graph 2102 is a relationshipbetween ε and IVC when the fuel is the low octane fuel.

If the geometric compression ratio ε is set as 10≤ε<17 in the engine 1of the high octane fuel, when the engine speed is 2,000 rpm, the designengineer determines the valve close timing IVC (deg.aBDC) so that thefollowing expression is satisfied.0.9949ε²−41.736ε+401.16≤IVC≤−0.9949ε²+41.736ε−361.16  (19)

If the geometric compression ratio ε is set as 17≤ε≤30 in the engine 1of the high octane fuel, when the engine speed is 2,000 rpm, the designengineer determines the valve close timing IVC (deg.aBDC) so that thefollowing expression is satisfied.−0.4288ε²+31.518ε−379.88≤IVC≤−0.9949ε²+41.736ε−361.16  (20)or0.9949ε²−41.736ε+401.16≤IVC≤1.9163ε²−89.935ε+974.94  (21)

By setting the valve close timing IVC of the intake valve 21 based onthe relational expressions (19) to (21), the SPCCI combustion of themixture gas with the A/F being leaner than the stoichiometric air fuelratio can be carried out, and the SPCCI combustion of the mixture gaswith the A/F being the stoichiometric air fuel ratio or richer than thestoichiometric air fuel ratio, and the G/F being leaner than thestoichiometric air fuel ratio can be carried out.

Note that the valve close timing IVC is set for each operating statewhich is determined by the load and the engine speed in Layer 2 andLayer 3.

As illustrated by a broken line, if the geometric compression ratio ε isset as 10≤ε<17 in the engine 1 of the high octane fuel, when the enginespeed is 3,000 rpm, the design engineer determines the valve closetiming IVC (deg.aBDC) so that the following expression is satisfied.0.9949ε²−41.736ε+403.16≤IVC≤−0.9949ε²+41.736ε−359.16  (19⁻¹)

If the geometric compression ratio ε is set as 17≤ε≤30 in the engine 1of the high octane fuel, when the engine speed is 3,000 rpm, the designengineer determines the valve close timing IVC (deg.aBDC) so that thefollowing expression is satisfied.−0.4288ε²+31.518ε−377.88≤IVC≤−0.9949ε²+41.736ε−359.16  (20⁻¹)or0.9949ε²−41.736ε+403.16≤IVC≤1.9163ε²−89.935ε+976.94  (21⁻¹)

Similarly, as illustrated by a one-dot chain line, if the geometriccompression ratio is set as 10≤ε<17 in the engine 1 of the high octanenumber fuel, when the engine speed is 4,000 rpm, the design engineerdetermines the valve close timing IVC (deg.aBDC) so that the followingexpression is satisfied.0.9949ε²−41.736ε+409.16≤IVC≤−0.9949ε²+41.736ε−353.16  (19⁻²)

If the geometric compression ratio ε is set as 17≤ε≤30 in the engine 1of the high octane number fuel, when the engine speed is 4,000 rpm, thedesign engineer determines the valve close timing IVC (deg.aBDC) so thatthe following expression is satisfied.−0.4288ε²+31.518ε−371.88≤IVC≤−0.9949ε²+41.736ε−353.16  (20⁻²)or0.9949ε²−41.736ε+409.16≤IVC≤1.9163ε²−89.935ε+982.94  (21⁻²)

If the correction term C according to the engine speed NE (rpm) of theengine 1 is used similar to the above, the relational expression of εand IVC in Layer 2 and Layer 3 can be expressed as follows. If thegeometric compression ratio ε is 10≤ε<17,0.9949ε²−41.736ε+401.16+C≤IVC≤−0.9949ε²+41.736ε−361.16+C  (19⁻³)If the geometric compression ratio ε is 17≤ε≤30,−0.4288ε²+31.518ε−379.88+C≤IVC≤−0.9949ε²+41.736ε−361.16+C  (20⁻³)or0.9949ε²−41.736ε+401.16+C≤IVC≤1.9163ε²−89.935ε+974.94+C  (21⁻³)

Here, if the geometric compression ratio ε determined to be lower than17, the design engineer can determine IVC based on the relationalexpression (19⁻³). Since the selection range of IVC is wide, a degree offreedom in the design becomes high.

Moreover, as illustrated in the lower graph 2102 of FIG. 21, if thegeometric compression ratio ε is set as 10≤ε<15.7 in the engine 1 of thelow octane fuel, when the engine speed is 2,000 rpm, the design engineerdetermines the valve close timing IVC (deg.aBDC) so that the followingexpression is satisfied.0.9949ε²−39.149ε+348.59≤IVC≤−0.9949ε²+39.149ε−308.59  (22)

Moreover, if the geometric compression ratio ε is set as 15.7≤ε≤30 inthe engine 1 of the low octane fuel, when the engine speed is 2,000 rpm,the design engineer determines the valve close timing IVC (deg.aBDC) sothat the following expression is satisfied.−0.5603ε²+34.859ε−377.22≤IVC≤−0.9949ε²+39.149ε−308.59  (23)or0.9949ε²−39.149ε+348.59≤IVC≤1.9211ε²−85.076ε+862.01  (24)

Although illustration is omitted, if the geometric compression ratio εis set as 10≤ε<15.7 in the engine 1 of the low octane fuel, when theengine speed is 3,000 rpm, the design engineer determines the valveclose timing IVC (deg.aBDC) so that the following expression issatisfied.0.9949ε²−39.149ε+350.59≤IVC≤−0.9949ε²+39.149ε−306.59  (22⁻¹)

Moreover, if the geometric compression ratio ε is set as 15.7≤ε≤30 inthe engine 1 of the low octane fuel, when the engine speed is 3,000 rpm,the design engineer determines the valve close timing IVC (deg.aBDC) sothat the following expression is satisfied.−0.5603ε²+34.859ε−375.22≤IVC≤−0.9949ε²+39.149ε−306.59  (23⁻¹)or0.9949ε²−39.149ε+350.59≤IVC≤1.9211ε²−85.076ε+864.01  (24⁻¹)

Although illustration is similarly omitted, if the geometric compressionratio ε is set as 10≤ε<15.7 in the engine 1 of the low octane fuel, whenthe engine speed is 4,000 rpm, the design engineer determines the valveclose timing IVC (deg.aBDC) so that the following expression issatisfied.0.9949ε²−39.149ε+356.59≤IVC≤−0.9949ε²+39.149ε−300.59  (22⁻²)

Moreover, if the geometric compression ratio ε is set as 15.7≤ε≤30 inthe engine 1 of the low octane fuel, when the engine speed is 4,000 rpm,the design engineer determines the valve close timing IVC (deg.aBDC) sothat the following expression is satisfied.−0.5603ε²+34.859ε−369.22≤IVC≤−0.9949ε²+39.149ε−300.59  (23⁻²)or0.9949ε²−39.149ε+356.59≤IVC≤1.9211ε²−85.076ε+870.01  (24⁻²)

If the correction term C according to the engine speed NE (rpm) of theengine 1 is used similar to the above, the relational expression of εand IVC in Layer 2 and Layer 3 in the engine 1 of the low octane fuelcan be expressed as follows. If the geometric compression ratio ε is10≤ε<15.7,0.9949ε²−39.149ε+348.59+C≤IVC≤−0.9949ε²+39.149ε−308.59+C  (22⁻³)If the geometric compression ratio ε is 15.7≤ε≤30,−0.5603ε²+34.859ε−377.22+C≤IVC≤−0.9949ε²+39.149ε−308.59+C  (23⁻³)or0.9949ε²−39.149ε+348.59+C≤IVC≤1.9211ε²−85.076ε+862.01+C  (24⁻³)

Note that although illustration is omitted, the design engineer maydetermine IVC within an overlapping range of the ε-IVC valid range ofthe upper graph 2101 and the ε-IVC valid range of the lower graph 2102of FIG. 21. Similar to the above, the design engineer can set thecontrol logic which suits both the engine 1 using the high octane fueland the engine 1 using the low octane fuel by determining IVC within theoverlapping range of the two occurring ranges.

(1-5) Procedure of Method of Implementing Control Logic

Next, a procedure of the method of implementing the control logic of theengine 1 for executing the SPCCI combustion will be described withreference to a flowchart illustrated in FIG. 22. The design engineer canexecute each step using a computer. The computer stores information onthe ε-IVC valid range illustrated in FIGS. 17, 20, and 21.

At Step S221 after the procedure starts, the design engineer first setsthe geometric compression ratio ε. The design engineer may input the setvalue of the geometric compression ratio ε into the computer.

At the following Step S222, the design engineer sets the valve openingangle of the intake valve 21, and the valve opening angle of the exhaustvalve 22. This corresponds to determining the cam shapes of the intakevalve 21 and the exhaust valve 22. The design engineer may input the setvalues of the valve opening angles of the intake valve 21 and theexhaust valve 22 into the computer. Thus, a hardware configuration ofthe engine 1 can be set at Steps S221 and S222.

At Step S223, the design engineer sets the operating state comprised ofthe load and the engine speed, and at the following Step S224, thedesign engineer selects IVC based on the ε-IVC valid range (FIGS. 17,20, and 21) stored in the computer.

Then, at Step S225, the computer determines whether the SPCCI combustioncan be achieved based on IVC set at Step S224. If the determination atStep S225 is YES, this procedure shifts to Step S226, and the designengineer determines the control logic of the engine 1 so that the SPCCIcombustion is performed in the operating state set at Step S223. On theother hand, if the determination at Step S225 is NO, this procedureshifts to Step S227, and the design engineer determines the controllogic of the engine 1 so that the SI combustion is performed in theoperating state set at Step S223. Note that at Step S227, the designengineer may again set the valve close timing IVC of the intake valve 21in consideration of performing the SI combustion.

As described above, the method of implementing the control logic of thecompression ignition engine disclosed herein determines the relationshipbetween the engine geometric compression ratio ε and the valve closetiming IVC of the intake valve 21. The design engineer can determine thevalve close timing IVC of the intake valve 21 within the range where therelationship is satisfied. The design engineer can implement the controllogic of the engine 1 by less time and labor compared with theconventional arts.

(2) Method of Implementing Control Logic According to Valve OpeningAngle of Intake Valve

As described above, the valve close timing IVC of the intake valve 21can be set based on the relationship between the geometric compressionratio ε and the valve close timing IVC of the intake valve 21.

The present inventors further found that, as a result of repeatingfurther diligent examinations, a given relationship was required betweenthe geometric compression ratio ε and the closed-valve period (valveopening angle) CA of the intake valve 21, in order to put to practicaluse the engine for executing the SPCCI combustion.

As described above, the engine 1 switches the layer between Layer 2 andLayer 3. Here, the relationship between the geometric compression ratioε and the valve opening angle CA of the intake valve 21 in Layer 2differs from the relationship between the geometric compression ratio εand the valve opening angle CA of the intake valve 21 in Layer 3.

(2-1) Relationship Between Geometric Compression Ratio and Valve OpeningAngle of Intake Valve in Layer 2

As described above, the engine 1 operates over the wide operating rangefrom the low load to the high load in Layer 2. In the operating range onthe higher load side, in order to achieve the SPCCI combustion in Layer2, the relationship between ε and IVC so that the temperature of thecombustion chamber 17 at θ_(CI) does not exceed the referencetemperature Tth1 is required.

Thus, the present inventors created a matrix of the valve close timingIVC of the intake valve 21 and the overlap period O/L of the intakevalve 21 and the exhaust valve 22, for every geometric compression ratioε (ε1, ε2 . . . ), and in this matrix, estimated the temperature of thecombustion chamber 17 during the CI combustion by using the model of theengine 1, while changing the values of IVC and O/L. When this estimatedresult becomes the reference temperature Tth1 or lower, the combinationof IVC and O/L can achieve the SPCCI combustion in Layer 2.

Thus, the present inventors obtained a relationship between thegeometric compression ratio ε, the valve close timing IVC of the intakevalve 21, and the overlap period O/L of the intake valve 21 and theexhaust valve 22, which can achieve the SPCCI combustion in Layer 2.According to this relation, when the combination of ε and IVC isdetermined, the range of O/L where the temperature becomes the referencetemperature Tth1 or lower can be determined. As already described, inLayer 2, the combination of ε and IVC which can perform the SPCCIcombustion within the operating range on the higher load side is asexpressed in the approximations (I) and (II) in FIG. 17.

That is, when the combination of ε and IVC which can perform the SPCCIcombustion in Layer 2 is selected based on FIG. 17, the range of O/Lcorresponding to this combination can be determined through thisrelation.

The magnitude of the overlap period O/L is determined based on the valveclose timing EVC of the exhaust valve 22 and the valve open timing IVOof the intake valve 21. Here, assumed that EVC and IVO were setsymmetrically with respect to an exhaust top dead center (symmetrical inthe direction of the crank angle), the present inventors obtained therange of IVO corresponding to the range of O/L.

For example, if O/L is 70 deg, EVC becomes 35 deg.aTDC and IVO becomes35 deg.bTDC. Then, the valve opening period of the intake valve 21,i.e., the valve opening angle CA, can be determined by using theobtained IVO, and IVC used for the determination of O/L.

In this way, the present inventors acquired the combination of ε and CAby converting the combination of ε and IVC where the temperature becomesthe reference temperature Tth1 or lower.

A graph 2301 in the upper graph of FIG. 23 illustrates theapproximations (A) and (B) calculated based on the combination of ε andCA. The horizontal axis of the graph 2301 is the geometric compressionratio ε, and the vertical axis is the valve opening angle CA (deg) ofthe intake valve 21. Although illustration is omitted the presentinventors plotted, on the graph 1701, the combination of ε and CA wherethe temperature becomes the reference temperature Tth1 or lower, anddetermined the approximations (A) and (B) based on the plots.

The graph 2301 corresponds to a case when the engine speed is 2,000 rpm.The approximations (A) and (B) are as follows.CA=−3.9394ε²+159.53ε−1314.9  Approximation (A):CA=0.9096ε²−47.634ε+745.28  Approximation (B):

In the graph 2301, when the combination of ε and CA is on the left sideof the approximations (A) and (B), the temperature of the combustionchamber 17 during the CI combustion becomes the reference temperatureTth1 or lower. This combination can carry out the SPCCI combustion ofthe mixture gas with the A/F being the stoichiometric air fuel ratio,and the G/F being leaner than the stoichiometric air fuel ratio.

The relationship between ε and CA described above is a relationshipbased on the maximum temperature of the combustion chamber 17 in Layer2.

Moreover, as described above, in Layer 2, in order to achieve the SPCCIcombustion when the engine 1 operates with the light load, therelationship between ε and IVC so that the temperature of the combustionchamber 17 at θ_(CI) becomes the reference temperature Tth2 or higher isrequired.

Thus, the present inventors created a matrix of the valve close timingIVC of the intake valve 21, and the overlap period O/L of the intakevalve 21 and the exhaust valve 22, for every geometric compression ratioε (ε1, ε2 . . . ), and in this matrix, estimated the temperature of thecombustion chamber 17 during CI combustion by using the model of theengine 1, while changing the values of IVC and O/L. The combination ofIVC and O/L where this estimated result becomes the referencetemperature Tth2 or higher can achieve the SPCCI combustion in Layer 2.

Thus, the present inventors obtained the relationship between thegeometric compression ratio ε, the valve close timing IVC of the intakevalve 21, and the overlap period O/L of the intake valve 21 and theexhaust valve 22, which can achieve the SPCCI combustion in Layer 2.According to this relation, when the combination of ε and IVC isdetermined, the range of O/L where the temperature becomes the referencetemperature Tth2 or higher can be determined. As already described, inLayer 2, the combination of ε and IVC which can perform the SPCCIcombustion within the operating range on the light-load side is asexpressed by the approximations (III) and (IV) in FIG. 17.

That is, when the combination of ε and IVC which can perform the SPCCIcombustion in Layer 2 is selected based on FIG. 17, the range of O/Lcorresponding to this combination can be determined through thisrelation.

The magnitude of the overlap period O/L is determined based on the valveclose timing EVC of the exhaust valve 22 and the valve open timing IVOof the intake valve 21. The present inventors obtained the range of IVOcorresponding to the range of O/L, by assuming that EVC and IVO were setsymmetrically with respect to an exhaust top dead center (symmetrical inthe direction of the crank angle). Thus, the valve opening period of theintake valve 21, i.e., the valve opening angle CA, can be determined byusing the obtained IVO, and IVC used for the determination of O/L.

In this way, the present inventors acquired the combination of ε and CAby converting the combination of ε and IVC where the temperature becomesthe reference temperature Tth2 or higher.

In the graph 2301 of the upper graph in FIG. 23, the approximations (C)and (D) calculated based on the combination of ε and CA where thetemperature becomes the reference temperature Tth2 or higher are alsoillustrated. The approximations (C) and (D) are as follows.CA=−0.4234ε²+22.926ε+42.164  Approximation (C):CA=0.4234ε²−22.926ε+417.84  Approximation (D):

In the graph 2301, when the combination of ε and CA is on the right sideof the approximations (C) and (D), the temperature of the combustionchamber 17 during the CI combustion becomes the reference temperatureTth2 or higher. This combination can carry out the SPCCI combustion ofthe mixture gas with the A/F being the stoichiometric air fuel ratio andthe G/F being leaner than the stoichiometric air fuel ratio.

As seen in FIG. 23, the relationship between ε and CA is substantiallyvertically symmetrical with respect to CA=about 230 deg. CA=230 degcorresponds to the valve close timing (the best IVC described above) atwhich IVC=20 deg.aBDC, i.e., the amount of gas introduced into thecombustion chamber 17 becomes the maximum, when the engine speed is2,000 rpm.

In FIG. 23, the combination of ε and CA within the range surrounded bythe approximations (A), (B), (C), and (D) is the combination which canput to practical use the engine 1 for executing the SPCCI combustion inLayer 2. In other words, other combinations of ε and CA outside thisrange cannot put to practical use the engine 1 for executing the SPCCIcombustion in Layer 2.

The design engineer must determine CA within the ε-CA valid range whichis hatched in FIG. 23, upon determining the valve opening angle CA ofthe intake valve 21 when the engine 1 operates in Layer 2.

For example, if the geometric compression ratio ε is set as 10≤ε<16, thedesign engineer determines the valve opening angle CA (deg) so that thefollowing expression is satisfied.0.4234ε²−22.926ε+417.84≤CA≤0.4234ε²+22.926ε+42.164  (25)

Moreover, if the geometric compression ratio ε is set as 16≤ε<20, thedesign engineer determines the valve opening angle CA (deg) so that thefollowing expression is satisfied.−3.9394ε²+159.53ε−1314.9≤CA≤−0.4234ε²+22.926ε+42.164  (26)or0.4234ε²−22.926ε+417.84≤CA≤−0.9096ε²−47.634ε+745.28  (27)

Further, if the geometric compression ratio ε is set as 20≤ε, the designengineer determines the valve opening angle CA (deg) so that thefollowing expression is satisfied.−3.9394ε²+159.53ε−1314.9≤CA  (28)orCA≤0.9096ε²−47.634ε+745.28  (29)

By setting the valve close timing IVC of the intake valve 21 based onthe relational expressions (25) to (29), the SPCCI combustion of themixture gas with the A/F being the stoichiometric air fuel ratio orricher than the stoichiometric air fuel ratio, and the G/F being leanerthan the stoichiometric air fuel ratio is achieved. Note that the valveclose timing IVC is set for each operating state which is determined bythe load and the engine speed in Layer 2. The example illustrated bysolid lines in FIG. 23 is the ε-CA valid range when the engine speed is2,000 rpm, as described above. If the engine speed changes, the ε-CAvalid range also changes. When the engine speed increases, the best IVCis retarded, and as a result, the corresponding CA is also retarded.

For example, when the engine speed is 3,000 rpm, CA corresponding to thebest IVC is about 232 deg. As illustrated by broken lines in FIG. 23,the ε-CA valid range when the engine speed is 3,000 rpm is obtained byparallelly translating the ε-CA valid range when the engine speed is2,000 rpm to the retard side by about 2 deg.

Therefore, if the geometric compression ratio ε is set as 10≤ε<16, whenthe engine speed is 3,000 rpm, the design engineer determines the valveopening angle CA (deg) so that the following expression is satisfied.0.4234ε²−22.926ε+419.84≤CA≤−0.4234ε²+22.926ε+44.164  (25′)

Moreover, if the geometric compression ratio ε is set as 16≤ε<20, thedesign engineer determines the valve opening angle CA (deg) so that thefollowing expression is satisfied.−3.9394ε²+159.53ε−1312.9≤CA≤−0.4234ε²+22.926ε+44.164  (26′)or0.4234ε²−22.926ε+419.84≤CA≤−0.9096ε²−47.634ε+747.28  (27′)

Further, if the geometric compression ratio ε is set as 20≤ε, the designengineer determines the valve opening angle CA (deg) so that thefollowing expression is satisfied.−3.9394ε²+159.53ε−1312.9≤CA  (28′)orCA≤0.9096ε²−47.634ε+747.28  (29′)

Moreover, when the engine speed is 4,000 rpm, CA corresponding to thebest IVC is about 238 deg. As illustrated by broken lines in FIG. 23,the ε-CA valid range when the engine speed is 3,000 rpm is obtained byparallelly translating the ε-CA valid range when the engine speed is2,000 rpm to the retard side by about 8 deg.

Therefore, if the geometric compression ratio ε is set as 10≤ε<16, whenthe engine speed is 4,000 rpm, the design engineer determines the valveopening angle CA (deg) so that the following expression is satisfied.0.4234ε²−22.926ε+425.84≤CA≤−0.4234ε²+22.926ε+50.164  (25″)

Moreover, if the geometric compression ratio ε is set as 16≤ε<20, thedesign engineer determines the valve opening angle CA (deg) so that thefollowing expression is satisfied.−3.9394ε²+159.53ε−1306.9≤CA≤−0.4234ε²+22.926ε+50.164  (26″)or0.4234ε²−22.926ε+425.84≤CA≤−0.9096ε²−47.634ε+753.28  (27″)

Further, if the geometric compression ratio ε is set as 20≤ε, the designengineer determines the valve opening angle CA (deg) so that thefollowing expression is satisfied.−3.9394ε²+159.53ε−1306.9≤CA  (28″)orCA≤0.9096ε²−47.634ε+753.28  (29″)

If a correction term D according to the engine speed NE (rpm) of theengine 1 is determined as D=3.3×10⁻¹NE³−1.0×10⁻⁶NE²+7.0×10⁻⁴NE, therelational expression of ε and CA in Layer 2 can be expressed asfollows. The valve opening angle CA (deg) is determined so that thefollowing expression is satisfied:

if the geometric compression ratio ε is 10≤ε<16,0.4234ε²−22.926ε+417.84+D≤CA≤−0.4234ε²+22.926ε+42.164+D   (25⁽³⁾),or if the geometric compression ratio ε is 16≤ε<20,−3.9394ε²+159.53ε−1314.9+D≤CA≤−0.4234ε²+22.926ε+42.164+D  (26⁽³⁾)or0.4234ε²−22.926ε+417.84+D≤CA≤0.9096ε²−47.634ε+745.28+D  (27⁽³⁾)

Further, if the design engineer determines the geometric compressionratio ε as 20≤ε,−3.9394ε²+159.53ε−1314.9+D≤CA  (28⁽³⁾)orCA≤0.9096ε²−47.634ε+745.28+D  (29⁽³⁾)

The design engineer determines the valve opening angle CA based on theε-CA valid range determined for every engine speed of the engine 1. As aresult, the design engineer can set the valve timing of the intake valve21 in Layer 2 as illustrated in FIG. 11.

(2-2) Change in ε-CA Valid Range by Difference of Octane Number

The graph 2301 of FIG. 23 is a relationship between ε and CA when thefuel is the high octane number fuel (octane number is about 96). A graph2302 illustrated in the lower graph is a relationship between ε and CAwhen the fuel is the low octane fuel (octane number is about 91).According to the examination by the present inventors, when the fuel wasthe low octane fuel, it was found that the ε-CA valid range shifts by1.3 compression ratios toward the lower compression ratio from the ε-CAvalid range of the high octane fuel.

Accordingly, upon determining the valve opening angle CA in the engine 1of the low octane fuel, if the geometric compression ratio ε is set as10≤ε<14.7, when the engine speed is 2,000 rpm, the design engineerdetermines the valve opening angle CA (deg) so that the followingexpression is satisfied.0.4234(ε+1.3)²−22.926(ε+1.3)+417.84≤CA≤−0.4234(ε+1.3)²+22.926(ε+1.3)+42.164  (30)

Moreover, if the geometric compression ratio ε is set as 14.7≤ε<18.7 inthe engine 1 of the low octane fuel, when the engine speed is 2,000 rpm,the design engineer determines the valve opening angle CA (deg) so thatthe following expression is satisfied.−3.9394(ε+1.3)²+159.53(ε+1.3)−1314.9≤CA≤−0.4234(ε+1.3)²+22.926(ε+1.3)+42.164  (31)or0.4234(ε+1.3)²−22.926(ε+1.3)+417.84≤CA≤−0.9096(ε+1.3)²−47.634(ε+1.3)+745.28  (32)

Further, if the geometric compression ratio ε is set as 18.7≤ε in theengine 1 of the low octane fuel, when the engine speed is 2,000 rpm, thedesign engineer determines the valve opening angle CA (deg) so that thefollowing expression is satisfied.−3.9394(ε+1.3)²+159.53(ε+1.3)−1314.9≤CA  (33)orCA≤0.9096(ε+1.3)²−47.634(ε+1.3)+745.28  (34)

The cross-hatched range in the graph 2302 of FIG. 23 is an overlappingrange of the ε-CA valid range of the high octane fuel and the ε-CA validrange of the low octane fuel. The design engineer can set the controllogic which suits both the engine 1 using the high octane fuel and theengine 1 using the low octane fuel by determining CA within theoverlapping range of the two occurring ranges. The design engineer cancollectively implement the engine control logic even if the octanenumber of fuel differs for every destination. The package design has anadvantage which lessens the design time and labor.

Note that although illustration is omitted, also in the engine 1 of thelow octane fuel, when the engine speed increases, the ε-CA valid rangeis parallelly translated to the retard side.

Accordingly, upon determining the valve opening angle CA in the engine 1of the low octane fuel, if the geometric compression ratio ε is set as10≤ε<14.7, when the engine speed is 3,000 rpm, the design engineerdetermines the valve opening angle CA (deg) so that the followingexpression is satisfied.0.4234(ε+1.3)²−22.926(ε+1.3)+419.84≤CA≤−0.4234(ε+1.3)²+22.926(ε+1.3)+44.164  (30′)

Moreover, if the geometric compression ratio ε is set as 14.7≤ε<18.7 inthe engine 1 of the low octane fuel, when the engine speed is 2,000 rpm,the design engineer determines the valve opening angle CA (deg) so thatthe following expression is satisfied.−3.9394(ε+1.3)²+159.53(ε+1.3)−1312.9≤CA≤−0.4234(ε+1.3)²+22.926(ε+1.3)+44.164  (31′)or0.4234(ε+1.3)²−22.926(ε+1.3)+419.84≤CA≤−0.9096(ε+1.3)²−47.634(ε+1.3)+747.28  (32′)

Further, upon determining the valve opening angle CA, if the geometriccompression ratio ε is set as 10≤ε<14.7 in the engine 1 of the lowoctane fuel, when the engine speed is 4,000 rpm, the design engineerdetermines the valve opening angle CA (deg) so that the followingexpression is satisfied.0.4234(ε+1.3)²−22.926(ε+1.3)+425.84≤CA≤−0.4234(ε+1.3)²+22.926(ε+1.3)+50.164  (30″)

Moreover, if the geometric compression ratio ε is set as 14.7≤ε<18.7 inthe engine 1 of the low octane fuel, when the engine speed is 4,000 rpm,the design engineer determines the valve opening angle CA (deg) so thatthe following expression is satisfied.−3.9394(ε+1.3)²+159.53(ε+1.3)−1306.9≤CA≤−0.4234(ε+1.3)²+22.926(ε+1.3)+50.164  (31″)or0.4234(ε+1.3)²−22.926(ε+1.3)+425.84≤CA≤−0.9096(ε+1.3)²−47.634(ε+1.3)+753.28  (32″)

Further, if the geometric compression ratio ε is set as 18.7≤ε in theengine 1 of the low octane fuel, when the engine speed is 4,000 rpm, thedesign engineer determines the valve opening angle CA (deg) so that thefollowing expression is satisfied.−3.9394(ε+1.3)²+159.53(ε+1.3)−1306.9≤CA  (33″)orCA≤0.9096(ε+1.3)²−47.634(ε+1.3)+753.28  (34″)

If the correction term D according to the engine speed NE (rpm) of theengine 1 is used similar to the above, the relational expression of εand CA in Layer 2 in the engine 1 of the low octane fuel can beexpressed as follows. If the geometric compression ratio ε is 10≤ε<14.7,0.4234(ε+1.3)²−22.926(ε+1.3)+417.84+D≤CA≤−0.4234(ε+1.3)²+22.926(ε+1.3)+42.164+D  (30⁽³⁾)If the geometric compression ratio ε is 14.7≤ε<18.7,−3.9394(ε+1.3)²+159.53(ε+1.3)−1314.9+D≤CA≤−0.4234(ε+1.3)²+22.926(ε+1.3)+42.164+D  (31⁽³⁾)or0.4234(ε+1.3)²−22.926(ε+1.3)+417.84+D≤CA≤0.9096(ε+1.3)²−47.634(ε+1.3)+745.28+D  (32⁽³⁾)If the geometric compression ratio ε is 18.7<6,−3.9394(ε+1.3)²+159.53(ε+1.3)−1314.9+D≤CA  (33⁽³⁾)orCA≤0.9096(ε+1.3)²−47.634(ε+1.3)+745.28+D  (34⁽³⁾)(2-3) Relationship Between Geometric Compression Ratio and Valve OpeningAngle of Intake Valve in Layer 3

As described above, regarding Layer 3, the state where the engine 1operates with the light load corresponds to the operating state at thelimit where the SPCCI combustion can occur. In the operating range onthe light-load side, in order to achieve the stable SPCCI combustion inLayer 3, the relationship between ε and IVC so that the temperatureinside the combustion chamber 17 at θ_(CI) becomes the referencetemperature Tth3 or higher is required.

Therefore, the present inventors performed the examination describedabove.

First, the present inventors estimated the temperature of the combustionchamber 17 during the CI combustion by using the model of the engine 1,while changing the values of IVC and O/L in the matrix (see FIG. 19) ofthe valve close timing IVC of the intake valve 21, and the overlapperiod O/L of the intake valve 21 and the exhaust valve 22, for everygeometric compression ratio ε (ε1, ε2 . . . ). The combination of IVCand O/L where this estimated result becomes the reference temperatureTth3 or lower can achieve the stable SPCCI combustion in Layer 3.

Then, the present inventors estimated the G/F by using the model of theengine 1 in the matrix comprised of the two parameters of IVC and O/L,in order to reduce the discharge of raw NO_(x). The combination of IVCand O/L where the G/F becomes the given value or higher can reduce thedischarge of raw NO_(x).

Then, the present inventors overlapped the combination of IVC and O/Lwhere the temperature becomes the reference temperature Tth3 or higherwith the combination of IVC and O/L where the G/F becomes the givenvalue or higher to determine the relationship between ε and IVC whichcan achieve both the stability of the SPCCI combustion and the reductionof raw NO_(x) discharged.

Here, the relationship between ε and IVC is based on the combination ofIVC and O/L. According to this relation, if the combination of ε and IVCis determined, the range of O/L where the temperature becomes thereference temperature Tth3 or lower and the G/F becomes the given valueor higher can be determined.

Moreover, as already described, in Layer 3, the combination of ε and IVCwhich can perform the SPCCI combustion within the operating range on thelight-load side is as expressed by the approximations (V) and (VI) inFIG. 20.

That is, if the combination of ε and IVC which can perform the SPCCIcombustion in Layer 3 is selected based on FIG. 20, the range of O/Lcorresponding to this combination can be determined through thisrelation.

The magnitude of the overlap period O/L is determined based on the valveclose timing EVC of the exhaust valve 22 and the valve open timing IVOof the intake valve 21. The present inventors obtained the range of IVOcorresponding to the range of O/L, by assuming that EVC and IVO were setsymmetrically with respect to an exhaust top dead center (symmetrical inthe direction of the crank angle). Thus, the valve opening period of theintake valve 21, i.e., the valve opening angle CA, can be determined byusing the obtained IVO, and IVC used for the determination of O/L.

In this way, the present inventors acquired the combination of ε and CA,by converting the combination of ε and IVC where the temperature becomesthe reference temperature Tth3 or lower and the G/F becomes the givenvalue or higher.

FIG. 24 illustrates the approximations (E), (F), (G), (H), (I), and (J)calculated from the combination of ε and CA. The horizontal axis in FIG.24 is the geometric compression ratio ε, and the vertical axis is thevalve opening angle CA (deg) of the intake valve 21.

An upper graph 2401 of FIG. 24 corresponds to a case where the enginespeed is 2,000 rpm. Both the approximations (E) and (F) are effectedwhen ε<14, and they are as follows.CA=60ε−550  Approximation (E):CA=−40ε+800  Approximation (F):

Moreover, both approximations (G) and (H) are effected when 14≤ε<16.3,and they are as follows.CA=290  Approximation (G):CA=−0.7246ε²+6.7391ε+287.68  Approximation (H):

Moreover, both the approximations (I) and (J) are effected when16.3≤ε<20, and they are as follows.CA=290  Approximation (I):CA=−12.162ε+403.24  Approximation (J):

In FIG. 24, the combination of ε and CA between a boundary connectingthe approximations (E), (G), and (I), and a boundary connecting theapproximations (F), (H), and (J) achieves the SPCCI combustion of themixture gas with the temperature of the combustion chamber 17 during theCI combustion becoming the reference temperature Tth3 or higher, and theA/F being leaner than the stoichiometric air fuel ratio.

The relationship between ε and CA described above is a relation, inLayer 3, based on the minimum temperature of the combustion chamber 17which can achieve the SPCCI combustion when the engine 1 operates withthe light load.

Moreover, the approximations (G) and (I) correspond to the retardinglimit of IVC and the maximum value of O/L, respectively.

On the other hand, regardless of Layer 2 and Layer 3, if the temperatureinside the combustion chamber 17 is too high, the CI combustion beginsbefore the start of the SI combustion, and thereby the SPCCI combustioncannot be performed.

Moreover, as described above, in Layer 3, if the mean temperature insidethe combustion chamber 17 at a compression top dead center reaches thereference temperature Tth4, it is thought that the CI combustion willbegin before the SI combustion begins, and in this case, the SPCCIcombustion cannot be performed.

Thus, the present inventors created a matrix of the valve close timingIVC of the intake valve 21 and the overlap period O/L of the intakevalve 21 and the exhaust valve 22, for every geometric compression ratioε (ε1, ε2 . . . ), and in this matrix, estimated the temperature of thecombustion chamber 17 at a compression top dead center by using themodel of the engine 1, while changing the values of IVC and O/L. Thecombination of IVC and O/L where this estimated result becomes thereference temperature Tth4 or lower can achieve the SPCCI combustion inLayer 3.

Thus, the present inventors obtained a relationship of the geometriccompression ratio ε, the valve close timing IVC of the intake valve 21,and the overlap period O/L of the intake valve 21 and the exhaust valve22, which can achieve the SPCCI combustion in Layer 3. According to thisrelation, if the combination of ε and IVC is determined, the range ofO/L where the temperature becomes the reference temperature Tth4 orlower can be determined. As already described, the combination of ε andIVC which can perform the SPCCI combustion in Layer 3 is as expressed bythe approximations (VII) and (VIII) in FIG. 20.

That is, when the combination of ε and IVC which can perform the SPCCIcombustion in Layer 3 is selected based on FIG. 20, the range of O/Lcorresponding to this combination can be determined through thisrelation.

The magnitude of the overlap period O/L is determined based on the valveclose timing EVC of the exhaust valve 22 and the valve open timing IVOof the intake valve 21. The present inventors obtained the range of IVOcorresponding to the range of O/L, by assuming that EVC and IVO were setsymmetrically with respect to an exhaust top dead center (symmetrical inthe direction of the crank angle). Thus, the valve opening period of theintake valve 21, i.e., the valve opening angle CA, can be determined byusing the obtained IVO, and IVC used for the determination of O/L.

In this way, the present inventors acquired the combination of ε and CAby converting the combination of ε and IVC where the temperature becomesthe reference temperature Tth4 or lower.

In the graph 2401 of the upper graph in FIG. 24, the approximations (M)and (N) calculated based on the combination of ε and CA where thetemperature becomes the reference temperature Tth4 or lower are alsoillustrated. The approximation (M) and (N) are as follows.CA=5ε+112.5  Approximation (M):CA=1.8571ε²−112.216+1834.6  Approximation (N):

Moreover, the approximations (M) and (N) are connected by theapproximations (O) and (P).CA=32.5ε−602.5  Approximation (O):ε=25  Approximation (P)

For example, the approximations (M) and (O) are connected at ε=26, andthe approximations (O) and (P), and the approximations (P) and (N) areconnected at ε=25.

In FIG. 24, the combination of ε and CA on the left side of theapproximations (M), (N), (O), and (P) can avoid that the CI combustionbegins before the SI combustion, thereby achieving the SPCCI combustion.

Moreover, in the graph 2401 of the upper graph in FIG. 24,approximations (K) and (L) which constitute the boundary connecting theapproximations (I) and (M), and the boundary connecting theapproximations (J) and (N) are also illustrated. The approximations (K)and (L) are as follows.CA=−4.1176ε+372.35  Approximation (K):CA=−1.7647ε+195.29  Approximation (L):

The approximation (K) corresponds to a retarding limit of the valveclose timing of the intake valve 21 set in consideration of the amountof gas introduced into the combustion chamber 17 when the engine 1operates in Layer 3, and the approximation (L) corresponds to anadvancing limit of the valve close timing of the intake valve 21 set inconsideration of the amount of gas introduced into the combustionchamber 17.

Upon determining the valve close timing IVC of the intake valve 21 whenthe engine 1 operates in Layer 3, the design engineer must determine CAwithin the ε-CA valid range (the hatched range in FIG. 24) surrounded bythe approximations (E), (F), (G), (H), (I), (K), (L), (M), (O), and (P)in FIG. 20.

For example, if the geometric compression ratio ε is set as 13.5≤ε<14,when the engine speed is 2,000 rpm, the design engineer determines thevalve opening angle CA (deg) so that the following expression issatisfied.−40ε+800≤CA≤60ε−550  (35)

If the geometric compression ratio ε is set as 14≤ε<16.3, when theengine speed is 2,000 rpm, the design engineer determines the valveopening angle CA (deg) so that the following expression is satisfied.−0.7246ε²+6.7391ε+287.68≤CA≤290  (36)

Further, if the geometric compression ratio ε is set as 16.3≤ε<20, whenthe engine speed is 2,000 rpm, the design engineer determines the valveopening angle CA (deg) so that the following expression is satisfied.−12.162ε+403.24≤CA≤290  (37)

Further, if the geometric compression ratio ε is set as 20≤ε<25, whenthe engine speed is 2,000 rpm, the design engineer determines the valveopening angle CA (deg) so that the following expression is satisfied.−1.7647ε+195.29≤CA≤−4.1176ε+372.35  (38)

Further, if the geometric compression ratio ε is set as 25≤ε≤28.5, whenthe engine speed is 2,000 rpm, the design engineer determines the valveopening angle CA (deg) so that the following expression is satisfied.5ε+112.5≤CA≤−4.1176ε+372.35  (39)or−1.7647ε+195.29≤CA≤1.8571ε²−112.21ε+1834.6  (40)

By setting the valve opening angle CA of the intake valve 21 based onthe relational expressions (35) to (40), the SPCCI combustion of themixture gas with the A/F being leaner than the stoichiometric air fuelratio is achieved. Note that the valve opening angle CA is set for eachoperating state which is determined by the load and the engine speed inLayer 3.

The example illustrated by the solid line in FIG. 24 is the ε-CA validrange when the engine speed is 2,000 rpm, as described above. If theengine speed changes, the ε-CA valid range also changes. If the enginespeed increases, the ε-CA valid range is also parallelly translated tothe retard side in FIG. 24.

For example, if the geometric compression ratio ε is set as 13.5≤ε<14,when the engine speed is 3,000 rpm (see a broken line), the designengineer determines the valve opening angle CA (deg) so that thefollowing expression is satisfied.−40ε+802≤CA≤60ε−548  (35′)

If the geometric compression ratio ε is set as 14≤ε<16.3, when theengine speed is 3,000 rpm, the design engineer determines the valveopening angle CA (deg) so that the following expression is satisfied.−0.7246ε²+6.7391ε+289.68≤CA≤292  (36′)

Further, if the geometric compression ratio ε is set as 16.3≤ε<20, whenthe engine speed is 3,000 rpm, the design engineer determines the valveopening angle CA (deg) so that the following expression is satisfied.−12.162ε+405.24≤CA≤292  (37′)

Further, if the geometric compression ratio ε is set as 20≤ε<25, whenthe engine speed is 3,000 rpm, the design engineer determines the valveopening angle CA (deg) so that the following expression is satisfied.−1.7647ε+197.29≤CA≤−4.1176ε+374.35  (38′)

Further, if the geometric compression ratio ε is set as 25≤ε≤28.5, whenthe engine speed is 3,000 rpm, the design engineer determines the valveopening angle CA (deg) so that the following expression is satisfied.5ε+114.5≤CA≤−4.1176ε+374.35  (39′)or−1.7647ε+197.29≤CA≤1.8571ε²−112.21ε+1836.6  (40′)

If the geometric compression ratio ε is set as 13.5≤ε<14, when theengine speed is 4,000 rpm (see a one-dot chain line), the designengineer determines the valve opening angle CA (deg) so that thefollowing expression is satisfied.−40ε+808≤CA≤60ε−542  (35″)

If the geometric compression ratio ε is set as 14≤ε<16.3, when theengine speed is 4,000 rpm, the design engineer determines the valveopening angle CA (deg) so that the following expression is satisfied.−0.7246ε²+6.7391ε+295.68≤CA≤298  (36″)

Further, if the geometric compression ratio ε is set as 16.3≤ε<20, whenthe engine speed is 4,000 rpm, the design engineer determines the valveopening angle CA (deg) so that the following expression is satisfied.−12.162ε+411.24≤CA≤298  (37″)

Further, if the geometric compression ratio ε is set as 20≤ε<25, whenthe engine speed is 4,000 rpm, the design engineer determines the valveopening angle CA (deg) so that the following expression is satisfied.−1.7647ε+203.29≤CA≤−4.1176ε+380.35  (38″)

Further, if the geometric compression ratio ε is set as 25≤ε≤28.5, whenthe engine speed is 4,000 rpm, the design engineer determines the valveopening angle CA (deg) so that the following expression is satisfied.5ε+120.5≤CA≤−4.1176ε+380.35  (39″)or−1.7647ε+203.29≤CA≤1.8571ε²−112.21ε+1842.6  (40″)

If the correction term D according to the engine speed NE (rpm) of theengine 1 is used similar to the above, the relational expression of εand CA in Layer 3 can be expressed as follows. If the geometriccompression ratio ε is 13.5≤ε<14,−40ε+800+D≤CA≤60ε−550+D  (35⁽³⁾)If the geometric compression ratio ε is 14≤ε<16.3,−0.7246ε²+6.7391ε+287.68+D≤CA≤290+D  (36⁽³⁾)If the geometric compression ratio ε is 16.3≤ε<20,−12.162ε+403.24+D≤CA≤290+D  (37⁽³⁾)If the geometric compression ratio ε is 20≤ε<25,−1.7647ε+195.29+D≤CA≤−4.1176ε+372.35+D  (38⁽³⁾)If the geometric compression ratio ε is 25≤ε≤28.5,5ε+112.5+D≤CA≤−4.1176ε+372.35+D  (39⁽³⁾)or−1.7647ε+195.29+D≤CA≤1.8571ε²−112.21ε+1834.6+D  (40⁽³⁾)

The design engineer determines the valve opening angle CA based on theε-CA valid range determined for every engine speed of the engine 1. As aresult, the design engineer can set the valve timing of the intake valve21 in Layer 3, as illustrated in FIG. 12.

Moreover, a lower graph 2402 of FIG. 24 is a relationship between ε andCA when the fuel is the low octane fuel.

Upon determining the valve opening angle CA, if the geometriccompression ratio ε is set as 12.2≤ε<12.7 in the engine 1 of the lowoctane fuel, when the engine speed is 2,000 rpm, the design engineerdetermines the valve opening angle CA (deg) so that the followingexpression is satisfied.−40(ε+1.3)+800≤CA≤60ε−550  (41)

If the geometric compression ratio ε is set as 12.7≤ε<15 in the engine 1of the low octane fuel, when the engine speed is 2,000 rpm, the designengineer determines the valve opening angle CA (deg) so that thefollowing expression is satisfied.−0.7246(ε+1.3)²+6.7391(ε+1.3)+287.68≤CA≤290  (42)

Further, if the geometric compression ratio ε is set as 15.3≤ε<18.7 inthe engine 1 of the low octane fuel, when the engine speed is 2,000 rpm,the design engineer determines the valve opening angle CA (deg) so thatthe following expression is satisfied.−12.162(ε+1.3)+403.24≤CA≤290  (43)

Further, if the geometric compression ratio ε is set as 18.7≤ε<23.7 inthe engine 1 of the low octane fuel, when the engine speed is 2,000 rpm,the design engineer determines the valve opening angle CA (deg) so thatthe following expression is satisfied.−1.7647(ε+1.3)+195.29≤CA≤−4.1176(ε+1.3)+372.35  (44)

Further, if the geometric compression ratio ε is set as 23.7≤ε<27.2 inthe engine 1 of the low octane fuel, when the engine speed is 2,000 rpm,the design engineer determines the valve opening angle CA (deg) so thatthe following expression is satisfied.5(ε+1.3)+112.5≤CA≤−4.1176(ε+1.3)+372.35  (45)or−1.7647(ε+1.3)+195.29≤CA≤1.8571(ε+1.3)²−112.21(ε+1.3)+1834.6  (46)

The cross-hatched range in the lower graph 2402 of FIG. 24 is anoverlapping range of the ε-CA valid range of the high octane fuel andthe ε-CA valid range of the low octane fuel. Similar to the above, thedesign engineer can set the control logic which suits both the engine 1using the high octane fuel and the engine 1 using the low octane fuel bydetermining CA within the overlapping range of the two occurring ranges.

Note that although illustration is omitted, also in the engine 1 of thelow octane fuel, when the engine speed increases, the ε-CA valid rangeis parallelly translated to the retard side.

Upon determining the valve opening angle CA, if the geometriccompression ratio ε is set as 12.2≤ε<12.7 in the engine 1 of the lowoctane fuel, when the engine speed is 3,000 rpm, the design engineerdetermines the valve opening angle CA (deg) so that the followingexpression is satisfied.−40(ε+1.3)+802≤CA≤60ε−548  (41′)

If the geometric compression ratio ε is set as 12.7≤ε<15 in the engine 1of the low octane fuel, when the engine speed is 3,000 rpm, the designengineer determines the valve opening angle CA (deg) so that thefollowing expression is satisfied.−0.7246(ε+1.3)²+6.7391(ε+1.3)+289.68≤CA≤292  (42′)

Further, if the geometric compression ratio ε is set as 15.3≤ε<18.7 inthe engine 1 of the low octane fuel, when the engine speed is 3,000 rpm,the design engineer determines the valve opening angle CA (deg) so thatthe following expression is satisfied.−12.162(ε+1.3)+405.24≤CA≤292  (43′)

Further, if the geometric compression ratio ε is set as 18.7≤ε<23.7 inthe engine 1 of the low octane fuel, when the engine speed is 3,000 rpm,the design engineer determines the valve opening angle CA (deg) so thatthe following expression is satisfied.−1.7647(ε+1.3)+197.29≤CA≤−4.1176(ε+1.3)+374.35  (44′)

Further, if the geometric compression ratio ε is set as 23.7≤ε≤27.2 inthe engine 1 of the low octane fuel, when the engine speed is 3,000 rpm,the design engineer determines the valve opening angle CA (deg) so thatthe following expression is satisfied.5(ε+1.3)+114.5≤CA≤−4.1176(ε+1.3)+374.35  (45′)or−1.7647(ε+1.3)+197.29≤CA≤1.8571(ε+1.3)²−112.21(ε+1.3)+1836.6  (46′)

If the geometric compression ratio ε is set as 12.7≤ε<15 in the engine 1of the low octane fuel, when the engine speed is 4,000 rpm, the designengineer determines the valve opening angle CA (deg) so that thefollowing expression is satisfied.−40(ε+1.3)+808≤CA≤60ε−542  (41″)

If the geometric compression ratio ε is set as 12.7≤ε<15 in the engine 1of the low octane fuel, when the engine speed is 4,000 rpm, the designengineer determines the valve opening angle CA (deg) so that thefollowing expression is satisfied.−0.7246(ε+1.3)²+6.7391(ε+1.3)+295.68≤CA≤298  (42″)

Further, if the geometric compression ratio ε is set as 15.3≤ε<18.7 inthe engine 1 of the low octane fuel, when the engine speed is 4,000 rpm,the design engineer determines the valve opening angle CA (deg) so thatthe following expression is satisfied.−12.162(ε+1.3)+411.24≤CA≤298  (43″)

Further, if the geometric compression ratio ε is set as 18.7≤ε<23.7 inthe engine 1 of the low octane fuel, when the engine speed is 4,000 rpm,the design engineer determines the valve opening angle CA (deg) so thatthe following expression is satisfied.−1.7647(ε+1.3)+203.29≤CA≤−4.1176(ε+1.3)+380.35  (44″)

Further, if the geometric compression ratio ε is set as 23.7≤ε≤27.2 inthe engine 1 of the low octane fuel, when the engine speed is 4,000 rpm,the design engineer determines the valve opening angle CA (deg) so thatthe following expression is satisfied.5(ε+1.3)+120.5≤CA≤−4.1176(ε+1.3)+380.35  (45″)or−1.7647(ε+1.3)+203.29≤CA≤1.8571(ε+1.3)²−112.21(ε+1.3)+1842.6  (46″)

If the correction term D according to the engine speed NE (rpm) of theengine 1 is used similar to the above, the relational expression of εand CA in Layer 3 in the engine 1 of the low octane fuel can beexpressed as follows. If the geometric compression ratio ε is12.2≤ε<12.7,−40(ε+1.3)+800+D≤CA≤60ε−550+D  (41⁽³⁾)If the geometric compression ratio ε is 12.7≤ε<15,−0.7246(ε+1.3)²+6.7391(ε+1.3)+287.68+D≤CA≤290+D  (42⁽³⁾)If the geometric compression ratio ε is 15.3≤ε<18.7,−12.162(ε+1.3)+403.24+D≤CA≤290+D  (43⁽³⁾)If the geometric compression ratio ε is 18.7≤ε<23.7,−1.7647(ε+1.3)+195.29+D≤CA≤−4.1176(ε+1.3)+372.35+D  (44⁽³⁾)If the geometric compression ratio ε is 23.7≤ε≤27.2,5(ε+1.3)+112.5+D≤CA≤−4.1176(ε+1.3)+372.35+D  (45⁽³⁾)or−1.7647(ε+1.3)+195.29+D≤CA≤1.8571(ε+1.3)²−112.21(ε+1.3)+1834.6+D  (46⁽³⁾)(2-4) Relationship Between Geometric Compression Ratio and Valve OpeningAngle of Intake Valve in Layers 2 and 3

FIG. 25 illustrates a relationship between the geometric compressionratio ε and the valve opening angle CA of the intake valve 21 where theSPCCI combustion is possible in both Layer 2 and Layer 3. Thisrelational expression is obtained from the ε-CA valid range of FIG. 23and the ε-CA valid range of FIG. 24.

When the ECU 10 selects Layer 3 according to the temperature, etc. ofthe engine 1, the low-load operating range of the engine 1 is switchedfrom Layer 2 to Layer 3. If the valve opening angle CA of the intakevalve 21 is set so that the SPCCI combustion is possible in both Layer 2and Layer 3, it becomes possible to continuously perform the SPCCIcombustion even when the map of the engine 1 is switched from Layer 2 toLayer 3.

An upper graph 2501 of FIG. 25 is a relationship between ε and CA whenthe fuel is the high octane fuel. A lower graph 2502 is a relationshipbetween ε and CA when the fuel is the low octane fuel.

If the geometric compression ratio ε is set as 13.5≤ε<14 in the engine 1of the high octane fuel, when the engine speed is 2,000 rpm, the designengineer determines the valve opening angle CA (deg) so that thefollowing expression is satisfied.−40ε+800≤CA≤60ε−550  (47)

If the geometric compression ratio ε is set as 14≤ε<16 in the engine 1of the high octane fuel, when the engine speed is 2,000 rpm, the designengineer determines the valve opening angle CA (deg) so that thefollowing expression is satisfied.−0.7246ε²+6.7391ε+287.68≤CA≤290  (48)

If the geometric compression ratio ε is set as 16≤ε<16.3 in the engine 1of the high octane fuel, when the engine speed is 2,000 rpm, the designengineer determines the valve opening angle CA (deg) so that thefollowing expression is satisfied.−3.9394ε²+159.53ε−1314.9≤CA≤290  (49)or−0.7246ε²+6.7391ε+287.68≤CA≤−0.9096ε²−47.634ε+745.28  (50)

If the geometric compression ratio ε is set as 16.3≤ε in the engine 1 ofthe high octane fuel, when the engine speed is 2,000 rpm, the designengineer determines the valve opening angle CA (deg) so that thefollowing expression is satisfied.−3.9394ε²+159.53ε−1314.9≤CA≤290  (51)or−12.162ε+403.24≤CA≤−0.9096ε²−47.634ε+745.28  (52)

By setting the valve opening angle CA of the intake valve 21 based onthe relational expressions (47) to (52), the SPCCI combustion of themixture gas with the A/F being leaner than the stoichiometric air fuelratio can be carried out, and the SPCCI combustion of the mixture gaswith the A/F being the stoichiometric air fuel ratio or richer than thestoichiometric air fuel ratio, and the G/F being leaner than thestoichiometric air fuel ratio can be carried out.

Note that the valve opening angle CA is set for each operating statewhich is determined by the load and the engine speed in Layer 2 andLayer 3.

As illustrated by a broken line, if the geometric compression ratio ε isset as 13.5≤ε<14 in the engine 1 of the high octane fuel, when theengine speed is 3,000 rpm, the design engineer determines the valveopening angle CA (deg) so that the following expression is satisfied.−40ε+802≤CA≤60ε−548  (47′)

If the geometric compression ratio ε is set as 14≤ε<16 in the engine 1of the high octane fuel, when the engine speed is 3,000 rpm, the designengineer determines the valve opening angle CA (deg) so that thefollowing expression is satisfied.−0.7246ε²+6.7391ε+289.68≤CA≤292  (48′)

If the geometric compression ratio ε is set as 16≤ε<16.3 in the engine 1of the high octane fuel, when the engine speed is 3,000 rpm, the designengineer determines the valve opening angle CA (deg) so that thefollowing expression is satisfied.−3.9394ε²+159.53ε−1312.9≤CA≤292  (49′)or−0.7246ε²+6.7391ε+289.68≤CA≤−0.9096ε²−47.634ε+747.28  (50′)

If the geometric compression ratio ε is set as 16.3≤ε in the engine 1 ofthe high octane fuel, when the engine speed is 3,000 rpm, the designengineer determines the valve opening angle CA (deg) so that thefollowing expression is satisfied.−3.9394ε²+159.53ε−1312.9≤CA≤292  (51′)or−12.162ε+405.24≤CA≤−0.9096ε²−47.634ε+747.28  (52′)

Similarly, as illustrated by a one-dot chain line, if the geometriccompression ratio is set as 13.5≤ε<14 in the engine 1 of the high octanefuel, when the engine speed is 4,000 rpm, the design engineer determinesthe valve opening angle CA (deg) so that the following expression issatisfied.−40ε+808≤CA≤60ε−542  (47″)

If the geometric compression ratio ε is set as 14≤ε<16 in the engine 1of the high octane fuel, when the engine speed is 4,000 rpm, the designengineer determines the valve opening angle CA (deg) so that thefollowing expression is satisfied.−0.7246ε²+6.7391ε+295.68≤CA≤298  (48″)

If the geometric compression ratio ε is set as 16≤ε<16.3 in the engine 1of the high octane fuel, when the engine speed is 4,000 rpm, the designengineer determines the valve opening angle CA (deg) so that thefollowing expression is satisfied.−3.9394ε²+159.53ε−1306.9≤CA≤298  (49″)or−0.7246ε²+6.7391ε+295.68≤CA≤−0.9096ε²−47.634ε+753.28  (50″)

If the geometric compression ratio ε is set as 16.3≤ε in the engine 1 ofthe high octane fuel, when the engine speed is 4,000 rpm, the designengineer determines the valve opening angle CA (deg) so that thefollowing expression is satisfied.−3.9394ε²+159.53ε−1306.9≤CA≤298  (51″)or−12.162ε+411.24≤CA≤−0.9096ε²−47.634ε+753.28  (52″)

If the correction term D according to the engine speed NE (rpm) of theengine 1 is used similar to the above, the relational expression of εand CA in Layer 2 and Layer 3 can be expressed as follows. If thegeometric compression ratio ε is 13.5≤ε<14,−40ε+800+D≤CA≤60ε−550+D  (47⁽³⁾)If the geometric compression ratio ε is 14≤ε<16,−0.7246ε²+6.7391ε+287.68+D≤CA≤290+D  (48⁽³⁾)If the geometric compression ratio ε is 16≤ε<16.3,−3.9394ε²+159.53ε−1314.9+D≤CA≤290+D  (49⁽³⁾)or−0.7246ε²+6.7391ε+287.68+D≤CA≤0.9096ε²−47.634ε+745.28+D  (50⁽³⁾)If the geometric compression ratio ε is 16.3≤ε,−3.9394ε²+159.53ε−1314.9+D≤CA≤290+D  (51⁽³⁾)or−12.162ε+403.24+D≤CA≤0.9096ε²−47.634ε+745.28+D  (52⁽³⁾)

Moreover, if the geometric compression ratio ε is set as 12.2≤ε<12.7 inthe engine 1 of the low octane fuel, when the engine speed is 2,000 rpm,the design engineer determines the valve opening angle CA (deg) so thatthe following expression is satisfied.−40(ε+1.3)+800≤CA≤60(ε+1.3)−550  (53)

If the geometric compression ratio ε is set as 12.7≤ε<14.7 in the engine1 of the low octane fuel, when the engine speed is 2,000 rpm, the designengineer determines the valve opening angle CA (deg) so that thefollowing expression is satisfied.−0.7246(ε+1.3)²+6.7391(ε+1.3)+287.68≤CA≤290  (54)

Moreover, if the geometric compression ratio ε is set as 14.7≤ε<15 inthe engine 1 of the low octane fuel, when the engine speed is 2,000 rpm,the design engineer determines the valve opening angle CA (deg) so thatthe following expression is satisfied.−3.9394(ε+1.3)²+159.53(ε+1.3)−1314.9≤CA≤290  (55)or−0.7246(ε+1.3)²+6.7391(ε+1.3)+287.68≤CA≤−0.9096(ε+1.3)²−47.634(ε+1.3)+745.28  (56)

If the geometric compression ratio ε is set as 15≤ε in the engine 1 ofthe low octane fuel, when the engine speed is 2,000 rpm, the designengineer determines the valve opening angle CA (deg) so that thefollowing expression is satisfied.−3.9394(ε+1.3)²+159.53(ε+1.3)−1314.9≤CA≤290  (57)or−12.162(ε+1.3)+403.24≤CA≤−0.9096(ε+1.3)²−47.634ε+745.28  (58)

Although illustration is omitted, if the geometric compression ratio εis set as 12.2≤ε<12.7 in the engine 1 of the low octane fuel, when theengine speed is 3,000 rpm, the design engineer determines the valveopening angle CA (deg) so that the following expression is satisfied.−40(ε+1.3)+802≤CA≤60(ε+1.3)−548  (53′)

If the geometric compression ratio ε is set as 12.7≤ε<14.7 in the engine1 of the low octane fuel, when the engine speed is 3,000 rpm, the designengineer determines the valve opening angle CA (deg) so that thefollowing expression is satisfied.−0.7246(ε+1.3)²+6.7391(ε+1.3)+289.68≤CA≤292  (54′)

Moreover, if the geometric compression ratio ε is set as 14.7≤ε<15 inthe engine 1 of the low octane fuel, when the engine speed is 3,000 rpm,the design engineer determines the valve opening angle CA (deg) so thatthe following expression is satisfied.−3.9394(ε+1.3)²+159.53(ε+1.3)−1312.9≤CA≤292  (55′)or−0.7246(ε+1.3)²+6.7391(ε+1.3)+289.68≤CA≤−0.9096(ε+1.3)²−47.634(ε+1.3)+747.28  (56′)

If the geometric compression ratio ε is set as 15≤ε in the engine 1 ofthe low octane fuel, when the engine speed is 3,000 rpm, the designengineer determines the valve opening angle CA (deg) so that thefollowing expression is satisfied.−3.9394(ε+1.3)²+159.53(ε+1.3)−1312.9≤CA≤292  (57′)or−12.162(ε+1.3)+405.24≤CA≤−0.9096(ε+1.3)²−47.634ε+747.28  (58′)

Although illustration is similarly omitted, if the geometric compressionratio ε is set as 12.2≤ε<12.7 in the engine 1 of the low octane fuel,when the engine speed is 4,000 rpm, the design engineer determines thevalve opening angle CA (deg) so that the following expression issatisfied.−40(ε+1.3)+808≤CA≤60(ε+1.3)−542  (53″)

Moreover, if the geometric compression ratio ε is set as 12.7≤ε<14.7 inthe engine 1 of the low octane fuel, when the engine speed is 4,000 rpm,the design engineer determines the valve opening angle CA (deg) so thatthe following expression is satisfied.−0.7246(ε+1.3)²+6.7391(ε+1.3)+295.68≤CA≤298  (54″)

Moreover, if the geometric compression ratio ε is set as 14.7≤ε<15 inthe engine 1 of the low octane fuel, when the engine speed is 4,000 rpm,the design engineer determines the valve opening angle CA (deg) so thatthe following expression is satisfied.−3.9394(ε+1.3)²+159.53(ε+1.3)−1306.9≤CA≤298  (55″)or−0.7246(ε+1.3)²+6.7391(ε+1.3)+295.68≤CA≤−0.9096(ε+1.3)²−47.634(ε+1.3)+753.28  (56″)

If the geometric compression ratio ε is set as 15≤ε in the engine 1 ofthe low octane fuel, when the engine speed is 4,000 rpm, the designengineer determines the valve opening angle CA (deg) so that thefollowing expression is satisfied.−3.9394(ε+1.3)²+159.53(ε+1.3)−1306.9≤CA≤298  (57″)or−12.162(ε+1.3)+411.24≤CA≤−0.9096(ε+1.3)²−47.634ε+753.28  (58″)

If the correction term D according to the engine speed NE (rpm) of theengine 1 is used similar to the above, the relational expression of εand CA in Layer 2 and Layer 3 in the engine 1 of the low octane fuel canbe expressed as follows.

If the geometric compression ratio ε is 12.2≤ε<12.7,−40(ε+1.3)+800+D≤CA≤60(ε+1.3)−550+D  (53⁽³⁾)If the geometric compression ratio ε is 12.7≤ε<14.7,−0.7246(ε+1.3)²+6.7391(ε+1.3)+287.68+D≤CA≤290+D  (54⁽³⁾)If the geometric compression ratio ε is 14.7≤ε<15,−3.9394(ε+1.3)²+159.53(ε+1.3)−1314.9+D≤CA≤290+D  (55⁽³⁾)or−0.7246(ε+1.3)²+6.7391(ε+1.3)+287.68+D≤CA≤0.9096(ε+1.3)²−47.634(ε+1.3)+745.28+D  (56⁽³⁾)If the geometric compression ratio ε is 15≤ε,−3.9394(ε+1.3)²+159.53(ε+1.3)−1314.9+D≤CA≤290+D  (57⁽³⁾)or−12.162(ε+1.3)+403.24+D≤CA≤0.9096(ε+1.3)²−47.634ε+745.28+D  (58⁽³⁾)

Note that although illustration is omitted, the design engineer maydetermine CA within the overlapping range of the ε-CA valid range of theupper graph 2501 and the ε-CA valid range of the lower graph 2502 ofFIG. 25. Similar to the above, the design engineer can set the controllogic which suits both the engine 1 using the high octane fuel and theengine 1 using the low octane fuel by determining CA within theoverlapping range of the two occurring ranges.

(2-5) Procedure of Method of Implementing Control Logic

Next, a procedure of the method of implementing the control logic of theengine 1 for executing the SPCCI combustion will be described withreference to the flowchart illustrated in FIG. 26. The design engineercan perform each step using a computer. The computer stores informationon the ε-CA valid range illustrated in FIGS. 23, 24, and 25.

At Step S261 after the procedure starts, the design engineer first setsthe geometric compression ratio ε. The design engineer may input the setvalue of the geometric compression ratio ε into the computer.

At the following Step S262, the design engineer sets the valve openingangle of the exhaust valve 22. This corresponds to determining the camshape of the exhaust valve 22. The design engineer may input the setvalue of the valve opening angle of the exhaust valve 22 into thecomputer. Thus, a hardware configuration of the engine 1 can be set atSteps S261 and S262.

At Step S263, the design engineer sets the operating state comprised ofthe load and the engine speed, and at the following Step S264, thedesign engineer selects CA based on the ε-CA valid range (FIGS. 23, 24,and 25) stored in the computer.

Then, at Step S265, the computer determines whether the SPCCI combustioncan be achieved based on CA set at Step S264. If the determination atStep S265 is YES, this procedure shifts to Step S266, and the designengineer determines the control logic of the engine 1 so that the SPCCIcombustion is performed in the operating state set at Step S263. On theother hand, if the determination at Step S265 is NO, this procedureshifts to Step S267, and the design engineer determines the controllogic of the engine 1 so that SI combustion is performed in theoperating state set at Step S263. Note that at Step S267, the designengineer may again set the valve opening angle CA of the intake valve 21in consideration of performing the SI combustion.

As described above, the method of implementing the control logic of thecompression ignition engine disclosed herein determines the relationshipbetween the engine geometric compression ratio ε and the valve openingangle CA of the intake valve 21. The design engineer can determine thevalve opening angle CA of the intake valve 21 within the range where therelationship is satisfied. The design engineer can implement the controllogic of the engine 1 by less time and labor compared with theconventional arts.

Other Embodiments

Note that the application of the technology disclosed herein is notlimited to the engine 1 having the configuration described above. Theengine 1 may adopt various configurations.

For example, the engine 1 may be provided with a turbocharger, insteadof the mechanical supercharger 44.

It should be understood that the embodiments herein are illustrative andnot restrictive, since the scope of the invention is defined by theappended claims rather than by the description preceding them, and allchanges that fall within metes and bounds of the claims, or equivalenceof such metes and bounds thereof, are therefore intended to be embracedby the claims.

DESCRIPTION OF REFERENCE CHARACTERS

1 Engine

10 ECU (Control Part)

17 Combustion Chamber

23 Intake-side Electric S-VT (Variable Valve Operating Mechanism)

25 Ignition Plug (Ignition Part)

44 Supercharger

55 EGR System

6 Injector (Fuel Injection Part)

SW1 Airflow Sensor

SW2 First Intake-air Temperature Sensor

SW3 First Pressure Sensor

SW4 Second Intake-air Temperature Sensor

SW5 Second Pressure Sensor

SW6 Pressure Indicating Sensor

SW7 Exhaust Temperature Sensor

SW8 Linear O₂ Sensor

SW9 Lambda O₂ Sensor

SW10 Water Temperature Sensor

SW11 Crank Angle Sensor

SW12 Accelerator Opening Sensor

SW13 Intake Cam Angle Sensor

SW14 Exhaust Cam Angle Sensor

SW15 EGR Pressure Difference Sensor

SW16 Fuel Pressure Sensor

SW17 Third Intake-air Temperature Sensor

What is claimed is:
 1. A method of implementing control logic of acompression ignition engine, the engine comprising: an injectorconfigured to inject fuel to be supplied in a combustion chamber; avariable valve operating mechanism configured to change a valve timingof an intake valve; an ignition plug configured to ignite a mixture gasinside the combustion chamber; at least one sensor configured to measurea parameter related to an operating state of the engine; and a processorconfigured to output a signal to the injector, the variable valveoperating mechanism, and the ignition plug, according to control logiccorresponding to the operating state of the engine, in response to themeasurement of the at least one sensor, the processor outputting thesignal to the ignition plug in a specific operating state defined by anengine load and an engine speed so that unburnt mixture gas combusts byself ignition after the ignition plug ignites the mixture gas inside thecombustion chamber, the method comprising the steps of: determining ageometric compression ratio c of the engine; and determining the controllogic defining a valve opening angle CA of the intake valve, utilized bythe processor to generate a signal to the injector, the variable valveoperating mechanism, and the ignition plug, wherein, when determiningthe control logic, the valve opening angle CA (deg) is determined sothat the following expression is satisfied: if the geometric compressionratio ε is ε<14,−40ε+800+D≤CA≤60ε−550+D and if the geometric compression ratio ε is14≤ε<16,−0.7246ε2+6.7391ε+287.68+D≤CA≤290+D and if the geometric compressionratio ε is 16≤ε<16.3,−3.9394ε2+159.53ε−1314.9+D≤CA≤290+Dor−0.7246ε2+6.7391ε+287.68+D≤CA≤0.9096ε2−47.634ε+745.28+D and if thegeometric compression ratio ε is 16.3≤ε,−3.9394ε2+159.53ε−1314.9+D≤CA≤290+Dor−12.162ε+403.24+D≤CA≤0.9096ε2−47.634ε+745.28+D, where D is a correctionterm according to the engine speed NE (rpm),D=3.3×10−10NE3−1.0×10−6NE2+7.0×10−4NE.
 2. The method of claim 1, whereina valve close timing of the intake valve changes as the operating stateof the engine changes, and wherein the valve opening angle CA (deg) isdetermined for each operating state.
 3. The method of claim 1, whereinthe engine operates in a low-load operating state where the load is agiven load or lower, according to the control logic.
 4. The method ofclaim 1, wherein the engine has: a first mode in which the processoroutputs a signal to each of the injector and the variable valveoperating mechanism so that an air-fuel ratio (A/F) that is a weightratio of air contained in the mixture gas inside the combustion chamberto the fuel becomes leaner than a stoichiometric air fuel ratio, andoutputs a signal to the ignition plug so that the unburnt mixture gascombusts by self ignition after the ignition plug ignites the mixturegas inside the combustion chamber, and a second mode in which theprocessor outputs a signal to each of the injector and the variablevalve operating mechanism so that a gas-fuel ratio (G/F) that is aweight ratio of the entire mixture gas inside the combustion chamber tothe fuel becomes leaner than the stoichiometric air fuel ratio, and theA/F becomes the stoichiometric air fuel ratio or richer than thestoichiometric air fuel ratio, and outputs a signal to the ignition plugso that the unburnt mixture gas combusts by self ignition after theignition plug ignites the mixture gas inside the combustion chamber. 5.The method of claim 1, wherein the engine is provided with an exhaustgas recirculation (EGR) system configured to introduce burnt gas intothe combustion chamber, and wherein the processor outputs a signal tothe EGR system and the ignition plug so that a heat amount ratio used asan index related to a ratio of an amount of heat generated when themixture gas combusts by flame propagation to the entire amount of heatgenerated when the mixture gas inside the combustion chamber combusts,becomes a target heat amount ratio defined corresponding to theoperating state of the engine.
 6. The method of claim 1, wherein theprocessor outputs a signal to the EGR system and the ignition plug sothat the heat amount ratio becomes higher when the load of the engine ishigher.
 7. A method of implementing control logic of a compressionignition engine using low octane fuel, the engine comprising: aninjector configured to inject fuel to be supplied in a combustionchamber; a variable valve operating mechanism configured to change avalve timing of an intake valve; an ignition plug configured to ignite amixture gas inside the combustion chamber; at least one sensorconfigured to measure a parameter related to an operating state of theengine; and a processor configured to output a signal to the injector,the variable valve operating mechanism, and the ignition plug, accordingto a control logic corresponding to the operating state of the engine,in response to the measurement of the at least one sensor, the processoroutputting the signal to the ignition plug in a specific operating statedefined by an engine load and an engine speed so that unburnt mixturegas combusts by self ignition after the ignition plug ignites themixture gas inside the combustion chamber, the method comprising thesteps of: determining that the fuel is a low octane fuel, determining ageometric compression ratio ε of the engine; and determining controllogic defining a valve opening angle CA of the intake valve, utilized bythe processor to generate a signal to the injector, the variable valveoperating mechanism, and the ignition plug, wherein, when determiningthe control logic, the valve opening angle CA (deg) is determined sothat the following expression is satisfied: if the geometric compressionratio ε is ε<12.7,−40(ε+1.3)+800+D≤CA≤60(ε+1.3)−550+D and if the geometric compressionratio ε is 12.7≤ε<14.7,−0.7246(ε+1.3)2+6.7391(ε+1.3)+287.68+D≤CA≤290+D and if the geometriccompression ratio ε is 14.7≤ε<15,−3.9394(ε+1.3)2+159.53(ε+1.3)−1314.9+D≤CA≤290+Dor−0.7246(ε+1.3)2+6.7391(ε+1.3)+287.68+D≤CA≤0.9096(ε+1.3)2−47.634(ε+1.3)+745.28+Dand if the geometric compression ratio ε is 15≤ε,−3.9394(ε+1.3)2+159.53(ε+1.3)−1314.9+D≤CA≤290+Dor−12.162(ε+1.3)+403.24+D≤CA≤0.9096(ε+1.3)2−47.634(ε+1.3)+745.28+D, whereD is a correction term according to the engine speed NE (rpm),D=3.3×10−10NE3−1.0×10−6NE2+7.0×10−4NE.
 8. The method of claim 7, whereina valve close timing of the intake valve changes as the operating stateof the engine changes, and wherein the valve opening angle CA (deg) isdetermined for each operating state.
 9. The method of claim 7, whereinthe engine operates in a low-load operating state where the load is agiven load or lower, according to the control logic.
 10. The method ofclaim 7, wherein the engine has: a first mode in which the processoroutputs a signal to each of the injector and the variable valveoperating mechanism so that an air-fuel ratio (A/F) that is a weightratio of air contained in the mixture gas inside the combustion chamberto the fuel becomes leaner than a stoichiometric air fuel ratio, andoutputs a signal to the ignition plug so that the unburnt mixture gascombusts by self ignition after the ignition plug ignites the mixturegas inside the combustion chamber, and a second mode in which theprocessor outputs a signal to each of the injector and the variablevalve operating mechanism so that a gas-fuel ratio (G/F) that is aweight ratio of the entire mixture gas inside the combustion chamber tothe fuel becomes leaner than the stoichiometric air fuel ratio, and theA/F becomes the stoichiometric air fuel ratio or richer than thestoichiometric air fuel ratio, and outputs a signal to the ignition plugso that the unburnt mixture gas combusts by self ignition after theignition plug ignites the mixture gas inside the combustion chamber. 11.The method of claim 7, wherein the engine is provided with an exhaustgas recirculation (EGR) system configured to introduce burnt gas intothe combustion chamber, and wherein the processor outputs a signal tothe EGR system and the ignition plug so that a heat amount ratio used asan index related to a ratio of an amount of heat generated when themixture gas combusts by flame propagation to the entire amount of heatgenerated when the mixture gas inside the combustion chamber combusts,becomes a target heat amount ratio defined corresponding to theoperating state of the engine.
 12. The method of claim 7, wherein theprocessor outputs a signal to the EGR system and the ignition plug sothat the heat amount ratio becomes higher when the load of the engine ishigher.